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JOURNAL OF DESIGN INNOVATION FOR HYDRONIC AND PLUMBING PROFESSIONALS
29
July 2021
filter
pool
Heat Exchangers in Hydronic
and Plumbing Systems
www.caleffi.com
CUTTING-EDGE
INNOVATION
IN TEMPERATURE MIXING
Caleffi mixing valves lead the way. From 3/8” under-sink scald protection valves to 3” flanged digital master mixing
valves, we have a full offering for residential and commercial applications. Over 50 years of innovation and
global experience assure high quality and proven reliability. A wide selection of double union connection types work
with copper, iron, steel and non-metallic pipes. The valves comply with the necessary standards and codes for the
U.S. and Canada. CALEFFI GUARANTEED.
FROM THE GENERAL MANAGER & CEO
Dear Plumbing and Hydronic Professional,
While relaxing with some friends, someone posed this hypothetical question to the group,
“If you awakened tomorrow to find the nation’s electrical grid had become disabled, what
would you do?” The discussion that followed made for interesting insight into the priorities
people set in life, as well as their trust (or distrust) of human
behavior. How long would food and water be available? How
would people manage to survive, and who would survive?
I suggest posing this question when you get together with
friends. The conversation that’s likely to follow is sure to be
interesting and entertaining.
Now substitute “heat exchangers” for “electrical grid” in
that hypothetical question. Would your friends arrive at
similar conclusions? Those who work with thermal systems
probably would. Why? Because without heat exchangers they
would realize that many of the devices we take for granted
simply wouldn’t function. Examples include refrigerators, furnaces, boilers, chillers, air
conditioners, dehumidifiers, computers, televisions, coffee makers, ovens, cell phones,
TVs, automobiles, and more. Practically every device that manages thermal energy
contains some type of heat exchanger. The same is true for electrical power plants —
which takes us right back to the original question about the electrical grid!
In this issue of idronics we’ll discuss how heat exchangers are used in heating, cooling
and plumbing applications. Different categories of heat exchangers are described, as are
common methods for assessing their performance. Installation details are provided that
ensure lasting peak performance. The issue concludes with several examples of how
heat exchangers can be applied within hydronic and plumbing systems. Some of these
applications are common, while others are unique.
We hope you enjoy this issue of idronics and encourage you to send us any feedback by
e-mailing us at idronics@caleffi.com.
An interactive edition of this issue is available at idronics.caleffi.com. The content is fully
searchable and is complimented with video and resource links. Prefer a hard copy? Visit us
at www.caleffi.us and click on the
icon to register.
Mark Olson
General Manager and CEO
3
TABLE TABLE OF OF CONTENTS
CONTENTS
SECTION PAGE
1 5 INTRODUCTION
5 Introduction
26 29
January July 2021
2020
7 Pipe & fitting nomenclature
Internal Pipe vs.
vs. tube
External heat
Pipe exchangers
size
Double Pipe fitting wall coil terminology heat exchangers
Liquid-to-liquid Biomass boilers heat exchangers
External Couplings heat exchangers
Shell & tube heat exchangers
Shell Elbows
& coil heat exchangers
2 8 HEAT EXCHANGER TYPES
Flat plate heat exchangers
Brazed plate heat exchangers
Plate & frame heat exchangers
Fan-forced liquid-to-air heat exchangers
Fan-coils
Air handlers
Greywater heat recovery heat exchangers
A Technical Journal
from
CALEFFI NORTH AMERICA, INC
3883 W. Milwaukee Rd
Milwaukee, Wisconsin 53208
USA
Tel: 414-238-2360
FAX: 414-238-2366
E-mail: idronics@caleffi.com
Website: www.caleffi.us
To receive future idronics issues
FREE, register online
www.caleffi.us
© Copyright 2019 2021
Caleffi North America, Inc.
Printed: Milwaukee, Wisconsin
USA
12 Metal pipe joining methods & materials
Conduction
Steel pipe
Convection
Stainless steel pipe
Wrought iron pipe
Black iron pipe
THERMAL Cast iron CHARACTERISTICS pipe
OF HEAT EXCHANGERS
Laminar Copper vs. water Turbulent tube flow
Reynolds Brass number
Flow Bronze direction through heat exchanger
Determining overall rate of heat
Metallurgical joining methods
transfer (lmtd method)
Overall Soft soldering rate of heat transfer –pipe
Brazing heat echanger
Welding
3 27 HEAT TRANSFER FUNDAMENTALS
4 33
5 51 INSTALLATION DETAILS FOR HEAT EXCHANGERS
78 APPENDIX A: CALEFFI HYDRONIC COMPONENTS
79 APPENDIX A: CALEFFI PLUMBING COMPONENTS
79 APPENDIX B: GENERIC COMPONENTS
Natural vs. forced convection
Thermal radiation
Determining overall rate of heat transfer
(effectiveness method)
Software-based heat exchanger selection
Heat exchanger performance indices
Thermal length of a heat exchanger
Thermal performance of water-to-air heat
exchangers
29 Polymer Heat exchanger pipe joining fouling
methods & materials Heat exchangers supplying
Chemical Polyvinyl fouling
chloride tubing (PVC)
antifreeze-protected circuits
Debris Chlorinated
fouling
PolyVinyl Chloride tubing (CPVC)
High-Density Polyethylene tubing (HDPE)
Cross-linked Polyethylene tubing (PEX)
Oxygen diffusion
Snow & ice melting (SIM) systems District heating
Composite PEX-AL-PEX tubing
Pool Heating
Heat interface units
Domestic Polyethylene-raised water heating
temperature tubing Lake (PE-RT) water coolings
On-demand PolyPropylene-random domestic
tubing (PP-R) Lake plate heat exchangers
Fusion water heating
joining methods
Summary
Heat exchangers for non-pressurized
thermal storage tanks
41 Summary
6 57 HEAT EXCHANGER APPLICATIONS
Disclaimer: Caleffi makes no warranty that the information presented in idronics meets the mechanical, electrical or other code requirements
applicable within a given jurisdiction. The diagrams presented in idronics are conceptual, and do not represent complete schematics for any
specific installation. Local codes may require differences in design, or safety devices relative to those shown in idronics. It is the responsibility
4
4 N° 55 giugno 2019
1. INTRODUCTION
One of humankind’s most significant achievements has
been the ability to generate heat and guide its movement
for useful purposes. This ability is essential to modern life,
especially when it’s used to establish and maintain thermal
comfort in habitable spaces. Heat exchangers of all types
are used for this purpose.
For centuries, heat was generated from fire, and methods
were developed to guide that heat into occupied areas of
buildings. An often-cited example of early heat exchange
for improving human comfort was the Roman hypocaust.
Wood-fueled fires would be maintained outside of
occupied spaces. The hot combustion gases from these
fires were channeled through spaces under stone floors
supported on stone columns, as seen in Figure 1-1, as
well as through hollow stone walls.
Today, modern heat exchangers allow heat transfer with
zero mixing of the liquid or gas supplying the heat with
the liquid or gas absorbing that heat.
Wood stoves are another example of centuries-old heat
exchangers that allow heat from combustion gases to be
transferred to occupied spaces with little, if any, release of
those combustion gases into those spaces.
Some wood and coal burning stoves also included
heat exchangers for heating domestic water. Figure 1-3
illustrates the concepts used in these early systems, which
Figure 1-2
Figure 1-1
Figure 1-3
hot
water
range
boiler
cold
water
water jacket
Because the combustion gases were significantly
warmer than the building surfaces, heat would be
absorbed into the stone. It would then transfer through
the stone by conduction and be released into occupied
spaces by thermal radiation and convection. This heat
exchange process took place with minimal mixing of the
combustion gases and the air in occupied spaces. As
such, the hypocaust functioned as a heat exchanger.
coal / wood!
burning!
water heater
5
were available long before electrically powered circulators.
Flow between the “water jacket” of the stove and the
“range boiler” tank was created by thermosiphoning due
to the differences in density between hot and cold water.
The evolution of heat exchangers was also critical to
development of internal combustion engines. Some of
the earliest engines were cooled solely by surrounding air.
As engine design improved and horsepower increased,
it became impractical to rely solely on surrounding air to
keep the engine temperature under control. Engineers of
that era turned to the superior thermal properties of water
as a means of conveying heat from inside engine blocks to
a location where it could be dissipated to surrounding air.
Figure 1-4 shows an example of a Ford Model T radiator.
radiators, this radiator was not pressurized. Model T
drivers learned to carry extra water with them to replace
the water lost through evaporation, and in some cases,
boiling inside the radiator.
All fuel-burning boilers used for heating buildings have a
combustion chamber combined with a heat exchanger.
Figure 1-5a shows an example of a cast iron section that
is used to build the heat exchanger of a cast iron boiler.
Figure 1-5b shows how this heat exchanger, which is also
called a boiler “block,” is made by joining several cast iron
sections together.
Figure 1-5
Figure 1-4
Courtesy of motormission.com
This radiator could be fundamentally described as a waterto-air
heat exchanger. For its time — the early 1900s —
it represented state-of-the-art-technology. Water from
the upper portion of the engine block flowed into the
upper portion of the radiator and divided up into multiple
closed channels made of copper or brass. Air passed
between these channels as a result of the car moving,
as well as flow created by a simple fan connected to the
engine’s crankshaft by a leather belt. The higher thermal
conductivity of the copper and brass channels provided
minimal thermal resistance between water and the outer
surfaces of the radiator. After giving up heat, the coolest
water settled into a reservoir at the base of the radiator
and flowed back to the lower portion of the engine. No
water pump was used. All flow was driven by the changes
in buoyancy of the water between the top of the engine
and lower portion of the radiator. Unlike modern vehicle
Hot gases pass upward from the combustion chamber
and across the “pins” on the cast iron sections. The pins
increase the heat transfer surface area of the section. Heat
from the hot gases passes through the cast iron walls of
each section and is absorbed by the water inside.
Heat exchanger technology continued to progress
through the 20th century. Hundreds of heat exchanger
designs were developed for use in boilers, radiators,
chillers, fan-coils and convectors. Wrought iron pipe and
copper tubing were embedded into concrete slabs to
create “radiant panel heat exchangers,” as seen in Figure
1-7. These panels transfer heat from heated water into
occupied spaces using thermal radiation and convection.
Today, heat exchangers are precisely engineered for use
in all types of stationary energy-processing equipment, as
well as virtually all land-based vehicles, marine vessels,
aircraft and spacecraft. These devices range from huge,
multi-ton cooling towers used to dissipate heat from highrise
buildings (Figure 1-8), to tiny liquid cooling systems for
microprocessors (Figure 1-9).
Courtesy of Weil McLain
6
Figure 1-6 Figure 1-8
Figure 1-9
Figure 1-7
Every hydronic heating and cooling system involves the
movement of thermal energy between water, or a waterbased
fluid, and one or more surrounding materials. This
heat exchange begins at a heat source and ends at one
or more heat emitters. With proper design, the rate of heat
exchange at all locations within the system allows for safe
operation and unsurpassed comfort.
Within hydronic heating and cooling systems, there are
many situations where it’s necessary to move heat from
one fluid to another without letting those fluids contact
each other. Examples include:
• Transferring heat from the “system” water in a boiler to
domestic water.
• Transferring heat from system water to an antifreeze
solution circulated through tubing for melting snow on
pavements.
• Transferring heat from an antifreeze solution circulating
through piping buried in the earth to the refrigerant
within a heat pump.
• Transferring heat from a district heating system to a
building located several miles away.
• Transferring heat from water circulating through a
cooling system to the refrigerant within a chiller.
This issue of idronics focuses on heat exchangers used
within hydronic heating and cooling systems, as well as
in some plumbing applications. The discussion ranges
from fundamental heat transfer concepts to performance
evaluation and novel applications within modern hydronic
systems.
7
2. HEAT EXCHANGER TYPES
A wide variety of heat exchangers
are used in modern hydronic heating
and cooling systems, as well as in
some plumbing applications. This
section illustrates the most common
hardware configurations and
describes common applications for
these heat exchangers.
INTERNAL vs. EXTERNAL
HEAT EXCHANGERS
In the context of hydronic systems,
internal heat exchangers are typically
helical coils made of a metal with high
thermal conductivity and suspended
within liquid-filled containers such
as a thermal storage tank. Figure
2-1 shows an example of an internal
coil heat exchanger permanently
mounted in the lower portion of a
pressure-rated stainless steel tank.
A heated fluid is circulated between
the heat source and the inside of
the coil heat exchanger. The coil is
surrounded by domestic water within
the tank’s shell. Heat transfers from
the outer coil surface to the domestic
water by natural convection.
There are several variations of tanks
with internal coil heat exchangers.
Courtesy of Heat-Flo, Inc.
Figure 2-1
Figure 2-2
1 internal coil
at bottom
Figure 2-2 shows some common
configurations.
Tanks with a single lower coil are
commonly used for domestic water
heating and are referred to as indirect
water heaters. Figure 2-3 shows
further details for this application.
The single lower coil releases heat
into the coolest water, which, due to
its density, collects at the bottom of
the tank. When hot domestic water is
drawn from the top of the tank, cold
domestic water enters through a dip
tube that carries it to the lower portion
of the tank. This helps to preserve
beneficial temperature stratification
within the tank.
The heat source could be a boiler,
heat pump or solar thermal collectors.
Hot fluid from the heat source enters
the upper coil connection and flows
downward, exiting at the lower
connection. This flow direction is
opposite the direction that water
within the tank flows after absorbing
heat from the coil. Because the two
fluids move in opposite directions,
this situation is called “counterflow”
heat exchange. Later sections will
show why counterflow is important
to maximize heat transfer rates.
1 internal coil
at top
Figure 2-3
to/from
heat source
2 internal coils
top and bottom
heated
domestic
water out
warm water
rising by
bouyancy
1 internal coil
at bottom
cold
domestic
water in
Some systems using heat pumps are
designed to supply the heat pump
with the coolest water drawn for the
lower portion of the tank. The lower
the water temperature at which the
heat pump operates, the higher its
efficiency. Figure 2-4 shows such a
system, but with a boiler supplying
supplemental (or backup) heat through
a coil heat exchanger mounted in the
upper portion of the tank.
8
Figure 2-4
supplemental / auxiliary boiler
heated
domestic
water
heat pump
water heater
domestic
water in
tank shell
1 internal coil
at top
cold
domestic
water
The upper coil heat exchanger allows the supplemental
heat to be added to ensure that water supplied from the
tank meets the desired setpoint, but without disturbing the
coolest water in the lower portion of the tank.
Another possibility is to route domestic water through a
coil heat exchanger in the upper portion of the tank, as
shown in Figure 2-5.
This requires the coil to be made from a material such as
copper or stainless steel that is compatible with domestic
water. The coil is surrounded by “system water” that
is heated by a heat source and is also used for space
heating. The tank in Figure 2-5 can also provide buffering
to a zoned distribution system.
Tanks having upper and lower coil heat exchangers can be
used for several applications. One is where the lower coil
provides heat input from a solar thermal collector array,
while the upper coil provides domestic water heating. The
water in the tank shell is heated by a boiler and provides
buffering for a highly zoned space-heating system. Figure
2-6 shows an example of such a system.
In some applications, the two internal coil heat exchangers
are connected in series to create a single heat exchanger
with increased surface area.
9
Figure 2-5
to / from load
internal coil
(copper or
stainless steel)
∆P valve
ASSE 1017
thermostatic
mixing valve
DHW
thermal
storage
tank
expansion
tank
cold
water
loading
unit
pellet boiler
CW
DOUBLE WALL COIL HEAT EXCHANGERS
Some mechanical codes require a special type of coil
heat exchanger when the application requires transferring
heat from an antifreeze solution to potable water. That
coil must be “double walled.” There must be a partial air
gap between the metal wall that receives heat from the
antifreeze solution and the metal wall that transfers heat
to potable water. This air gap creates a “leakage path”
that would allow a leak in either metal wall to exit the heat
exchanger outside of the tank, and thus provide visible
evidence of the leak. Figure 2-7 illustrates the concept.
Double wall heat exchangers create lower heat transfer
rates compared to single wall coils of equivalent surface
area. They also tend to have higher pressure drops at a
given flow rate relative to single wall heat exchangers.
Opinions vary on their use, as do mechanical code
requirements on when and where they must be used.
Internal coil heat exchangers, typically made of copper
tubing, are also used for heat input and heat extraction
from large non-pressurized thermal storage tanks often
used in systems supplied by cordwood gasification
or pellet-fired boilers. Figure 2-8 shows a typical
configuration.
In some systems, one suspended coil is used for heat
input while another is used for heat extraction. This is
the scenario shown in Figure 2-8. Notice that the flow
direction in each coil is opposite the direction in which
10
Figure 2-6
solar thermal
collector array
variable-speed
pressure-regulated
circulator
CW
DHW
zone
valves
ASSE
1017
mixer
load
load
load
load
temperature
sensor
boiler
tank water moves by natural convection. Flow inside
the heat input coil is from top to bottom. Flow inside
the heat extraction coil is from bottom to top. These
counterflow directions improve the rate of heat transfer
through each coil.
Piping from each coil passes through the tank’s sidewall
above the highest water level. The piping penetrations
are sealed to limit evaporation water loss but are typically
not rated for submerged operation.
The use of two coil heat exchangers introduces two
“thermal penalties” (e.g., undesirable but necessary
temperature drops) into the heat transfer process between
the heat source and the load. One is the temperature
drop between the water entering the heat input coil from
the heat source, and the average water temperature in
the tank. The other is the temperature drop between the
average tank temperature and the outlet temperature from
the heat extraction coil. The extent of these temperature
drops depends on the surface area of the coil and the
11
Figure 2-7
fluid 1
leakage path opening
outer wall
leakage path
inner wall
tank jacket
tank insulation
pressure vessel wall
fluid 1
fluid 2
helical coil double
wall heat exchanger
leakage path opening
Courtesy of HydroFlex Systems
Figure 2-8
heat
input
air
space
warmed
water
rising
cooled
water
descending
vent
heat
extraction
convection conditions between the
outer surface of each coil and the
tank water. In general, it’s desirable
to reduce or eliminate any thermal
penalty between the heat source and
the load. This is especially true for
heat sources that require relative low
operating temperatures to attain high
efficiency.
In some applications, multiple
submerged coil heat exchangers
are used to increase the total heat
transfer area. Figure 2-9 shows an
example of multiple copper coil heat
exchangers connected in parallel to
header piping.
Another approach is to “co-wind”
multiple copper tubes into a single
helical coil, as shown in Figure 2-10.
This provides additional surface area
while greatly reducing pressure drop
relative to a single coil of the same
total surface area. The individual coils
12
Courtesy of Hydroflex Systems, Inc.
Figure 2-9
Figure 2-10
water would be the source liquid, and an antifreeze solution
would be the destination liquid. In a cooling application, the
first word in the description would be the source of the
cooling effect, and the last word would be the destination
of the cooling effect.
LIQUID-TO-LIQUID HEAT EXCHANGERS
The three common types of liquid-to-liquid heat exchangers
used in hydronic systems are:
• Shell & tube heat exchangers
• Shell & coil heat exchangers
• Flat plate heat exchangers
Figure 2-11 illustrates the fundamental construction of
these heat exchangers.
Figure 2-11
Common heat exchanger types
upper
manifold
Courtesy of American Solartechnics
are manifolded together at each end of the heat exchanger,
allowing for a single supply and return connection.
EXTERNAL HEAT EXCHANGERS
The term “external” heat exchanger refers to any heat
exchanger that is not immersed in a fluid inside another
component such as a tank. Unless the term “internal” is
used to describe the heat exchanger, assume that it is
external, and thus just referred to as a heat exchanger.
The most commonly used heat exchangers in hydronic
systems can be categorized as follows:
• liquid-to-liquid
• liquid-to-air
The first word in these descriptions refers to the fluid
supplying heat. The final word refers to the fluid absorbing
heat. For example, in some snowmelting systems, boiler
shell
tubes
lower
manifold
shell
coil
side A
side B
shell & tube shell & coil flat plate
SHELL & TUBE HEAT EXCHANGERS
This heat exchanger design stems from the fundamental
concept of a pipe that separates two liquids exchanging
heat. One liquid surrounds the outer surface of the pipe
and the other flows inside the pipe. The fluid surrounding
the pipe is contained within another cylindrical vessel called
the shell, as shown in Figure 2-12.
In most shell & tube heat exchangers, the cooler fluid
passes through the annular space between the outer
surface of the tube and the shell. This reduces heat loss
from the outer surface of the shell.
Although it’s possible to create a shell & tube heat
exchanger with a single tube, as shown in Figure 2-12, the
length of the heat exchanger required to create sufficient
13
Figure 2-12
heat transfer
shell
tube
Figure 2-14
upper
manifold
shell
liquid
tube
liquid
shell
tubes
Figure 2-13
lower
manifold
Courtesy of Packless Industries
heat transfer area for many applications is excessive. In
some situations, this length issue is solved by coiling the
overall heat exchanger, as shown in Figure 2-13.
This component is called a coaxial tube-in-tube heat
exchanger. It is commonly used for water-to-refrigerant
heat transfer in heat pumps.
When coiling is not possible, multiple straight tubes are
used, as shown in Figure 2-14.
cross section
The tubes are welded or brazed to bulkhead plates near
each end of the heat exchanger. The two ends of the
exchanger serve as manifolds to distribute flow through the
multiple tubes. The second liquid passes through the shell
of the heat exchanger and around the outer surfaces of all
the tubes. These designs can be scaled up to products
capable of transferring several millions of Btu/hr.
Figure 2-15 shows an example of a large “tube bundle”
that would be inserted into an appropriately sized shell.
The baffle plates through which the tubes pass increase
turbulence within the shell, which increases the convective
heat transfer rate.
The tubes and shell are often made of different materials,
depending on the chemical nature of the fluids exchanging
heat, as well as the temperature and pressure at which the
heat exchanger is rated to operate.
14
Figure 2-15
Figure 2-16
Courtesy of Bell & Gossett
Shell & tube heat exchangers are commonly used for
industrial applications, often at high temperatures and
pressures. Many of these heat exchangers can be opened
for servicing or replacement of the tube bundle. They are
seldom used in smaller residential and light commercial
hydronic systems. Their design requires significantly more
metal volume relative to other heat exchanger designs of
comparable capacity. The physical size of a shell & tube
heat exchanger required to provide for a given internal area
is also larger than that of other heat exchanger designs.
Shell & tube heat exchangers also tend to have greater
external surface area relative to other designs and thus
experience higher parasitic heat loss for a given set of
operating conditions. They are also more prone to fouling
deposits due to lower internal flow velocities in the shell.
SHELL & COIL HEAT EXCHANGERS
A shell & coil heat exchanger is made by enclosing a
helically shaped coil of tubing within a metal or composite
shell. One fluid passes through the coil while the other
passes through the shell and over the outer surface of the
coil. Figure 2-17 shows two variations of shell & coil heat
exchangers.
The geometry of the shell varies. Some shells are designed
for vertical installation, with connection at the top and
bottom. Others are designed for installation in a variety of
orientations, with side connections on the shell and coil
connections at one end. Some shell & coil heat exchangers
have the coil permanently welded or brazed to the shell.
Courtesy of Bell & Gossett
Figure 2-17
Shell & tube heat exchangers are also available with multiple
tube passes. One liquid passes through half of the tube
bundle in one direction, reverses through U-bends at the
other end of the shell, and passes through the remaining
tubes, eventually returning to the same end of the heat
exchanger. Figure 2-16 shows an example of a 2-pass
shell & tube heat exchanger having a welded steel shell and
copper tubes. This heat exchanger has a removable cast
iron end cap that allows the tube bundle to be removed for
cleaning or replacement.
The concept of multiple tube passes can be increased
to 3-pass and 4-pass designs. Most shell & tube heat
exchangers also have internal baffle plates that increase
turbulence on the shell side for improved convective
heat transfer.
sealed shell & coil
heat exchanger
serviceable shell & coil
heat exchanger
15
Others allow the coil to be removed
by unbolting a sealed end plate at one
end of the shell.
Shell & coil heat exchangers have
been used for heat transfer between
two liquids as well as between a liquid
and a refrigerant. In the latter case, the
refrigerant typically passes through a
welded steel shell, while the liquid
passes through a copper coil. In some
systems, the volume of the shell also
serves as a liquid accumulator within a
refrigeration circuit.
Shell & coil heat exchangers are not
commonly used for liquid-to-liquid
heat transfer in hydronic systems.
One limitation is the amount of coil
surface area versus the overall size
of the heat exchanger. Another is
Figure 2-18
hot
domestic
water
air vent
w/ check
the higher fluid volume in the shell,
which increases thermal mass and
decreases the heat exchanger’s
response time to temperature
changes relative to that of other heat
exchanger designs.
One application where the increased
thermal mass of a shell & coil design is
desirable is when the heat exchanger
also serves as a thermal storage
device. A “reverse” indirect water
heater, as shown in Figure 2-18, is a
good example.
One can think of this reverse indirect
water heater as a high surface area
shell & coil heat exchanger with
added thermal mass and insulation.
Domestic water is fully heated on a
single upward pass through multiple
T&PRV
copper coils that are manifolded
together at the top and bottom of the
tank. Hot water from a boiler or other
heat source passes through the steel
shell of the tank, transferring heat to
the copper coils.
FLAT PLATE HEAT EXCHANGERS
One of the most contemporary devices
for fluid-to-fluid heat exchange is
called a flat plate heat exchanger. This
type of heat exchanger is now used
in many types of hydronic heating and
cooling systems, as well as for the
evaporator and condenser in some
refrigeration systems.
Fundamentally, a flat plate heat
exchanger is created by stacking
several pre-formed metal plates and
sealing the perimeter of those plates
together. The plates are shaped to
create narrow flow channels between
them. One fluid passes from one end
of the heat exchanger to the other
through the odd-numbered channels
(1, 3, 5, 6, etc.). The other fluid passes
from one end of the heat exchanger to
the other through the even-numbered
channels (2, 4, 6, 8, etc.). This concept
is illustrated in Figure 2-19.
hot water from
heat source
domestic water
passes up
through multiple
copper coils
Figure 2-20 shows examples of the
preformed stainless steel plates and
a partially disassembled flat plate
heat exchanger.
Figure 2-19
water back to
heat source
cold
domestic
water
16
Courtesy of (a) Alfa Laval, and (b) GEA PHE Systems
Figure 2-20a
Figure 2-20b
are 3 x 8 inches, 5 x 12 inches and 10 x 20 inches. For
a given plate size, the heat transfer capacity is increased
by adding plates to the “stack” that becomes the overall
heat exchanger. For example, a 5 x 12 x 40 flat plate heat
exchanger is built using 40 plates of nominal dimension
5 inches by 12 inches. One unique plate forms the back
of the heat exchanger (e.g., it has no holes through it).
Another unique plate forms the front of the heat exchanger
and transitions to the four piping connections.
Brazed plate heat exchangers are typically used in smallto
medium-scale hydronic systems, where heat transfer
rates up to several hundred thousand Btu/hr are required.
A small brazed plate heat exchanger might have 10 plates
measuring a nominal 3 inches by 8 inches. A large brazed
plate exchanger might have 100 plates measuring a
nominal 10 inches by 20 inches.
The placement of the holes within the plates creates four
internal manifold cavities. These cavities ensure even
distribution of both fluids into their respective channels.
The cavities transition to piping connections at one end of
the heat exchanger, as seen in Figure 2-20b.
The patterned plates are designed to encourage turbulent
flow, and thus high convection coefficients. Some
manufacturers offer different plate patterns to create internal
flows that are well-matched to specific fluid properties and
a desired flow rate and pressure drop characteristics.
For a given set of entering temperatures, flow rates and
fluid characteristics, the heat transfer capacity of a flat plate
heat exchanger depends primarily on the size and number
of plates used.
Flat plate heat exchangers can be further classified as:
Figure 2-21 shows three examples of brazed plate heat
exchangers: A small 3 x 8-inch x 10 plate, a 5 x 12-inch
x 30 plate, and a 5 x12-inch x 100 plate. The largest heat
exchanger has 1-1/4-inch MPT connections. The mediumsized
exchanger has 1 inch MPT connections. The small
heat exchanger has 3/4-inch FPT connections.
Most of the brazed plate heat exchangers used in hydronic
heating and cooling systems are constructed of 316
stainless steel. This allows them to operate with potable
water having some chloride content, as well as a range
of system fluids, including most antifreeze solutions. Some
manufacturers also offer heat exchangers made of lessexpensive
304 stainless steel for applications using fluids
with low chloride content. Brazed plate heat exchangers
typically have pressure and temperature ratings well above
those needed in hydronic systems.
Figure 2-21
• Brazed plate heat exchangers
• Plate & frame heat exchangers
BRAZED PLATE HEAT EXCHANGERS
As the name implies, all the plates on a brazed plate heat
exchanger are brazed together at their perimeter, as well
as where internal surfaces come in contact. The brazing is
typically done with a copper alloy. This metallurgically seals
the flow channels and creates two separate pressurerated
compartments within the heat exchanger. Once it is
brazed, the heat exchanger cannot be modified or opened.
Most manufacturers have standardized plates sizes for their
brazed plate heat exchangers. Typical plate dimensions
17
Figure 2-22
Heat exchangers with 5-inch x 12-inch (or larger)
plates should be supported by some type of bracket
to reduce stress on the connected piping. Figure 2-23a
shows an example of a fabricated bracket supporting an
insulated 5 x 12-inch x 100 plate heat exchanger. Figure
2-23b shows a 5 x 12-inch x 40 plate heat exchanger
supported on a small steel angle bracket screwed to a
plywood wall. Some brazed plate heat exchangers are
supplied with threaded studs that can be fastened to a
metal bracket, as shown in Figure 2-23c.
Brazed plate heat exchangers are also available with
double wall construction for situations where codes
require them. These units provide a leakage path
between adjacent plates that would route any leaked
fluid outside the heat exchanger.
Small brazed plate heat exchangers are relatively light.
A 3 x 8-inch heat exchanger can be supported by the 4
pipes connected to it. However, those pipes should be
supported within a few inches of the heat exchanger, as
shown in Figure 2-22.
PLATE & FRAME HEAT EXCHANGERS
Plate & frame heat exchangers could be considered the
“big brother” to brazed plate heat exchangers. They
use the same concept of a stack of preformed plates to
separate the two fluids in alternating channels. However,
instead of brazing, plate & frame heat exchangers use
gaskets to seal the fluid channels. The stack of plates is
assembled on a frame between two thick steel pressure
plates. When the required number of plates have been
loaded onto the frame, several threaded steel rods are
used to pull the pressure plates together and compress
the stack into a pressure-tight assembly. The plate stack,
pressure plates and tension rods can be seen on the heat
exchanger in Figure 2-24.
Figure 2-23a Figure 2-23b Figure 2-23c
18
Figure 2-24
Courtesy of Alfa Laval
Large plate & frame heat exchangers
are typically assembled close to, or
exactly at, their final position within a
mechanical room. Each plate is “hung”
from an upper rail and slid into initial
position to form the plate stack. The
threaded steel draw rods are carefully
tensioned using torque wrenches
to ensure even compression of the
stack. This assembly concept is
shown in Figure 2-25.
Figure 2-26 shows an exploded view
of a plate & frame heat exchanger. The
alternating fluid channels, thick steel
pressure plates and perimeter gaskets
are visible. Some plate & frame heat
exchangers also include a “shroud”
around the completed plate stack.
Large plate & frame heat exchangers
are heavy. They are usually anchored
to concrete pads within mechanical
rooms, as seen in Figure 2-27.
The large plate & frame heat
exchanger in Figure 2-27 has been
fully insulated to reduce heat loss into
Figure 2-25
Courtesy of Alfa Laval
Courtesy of Alfa Laval
Figure 2-26
19
Figure 2-27 Figure 2-28
the mechanical room. The welded
steel piping connected to this heat
exchanger has also been insulated
and jacketed. Note the installation of
thermometers near all four ports of
the heat exchanger. Drain valves that
tee into the two lower pipes can also
be seen.
The frames used for plate & frame
heat exchangers are usually large
enough to accommodate more plates
than the initial design requires. More
plates can be added if the capacity
of the original system is increased.
Figure 2-28 shows an example of the
additional “rack space” available on a
large plate & frame heat exchanger.
The extended ends of the tension
rods are typically covered with plastic
tubes for safety.
Although plate & frame heat
exchangers can be disassembled for
cleaning or plate replacement, this is
a laborious task. It is always better
to use proper dirt separation details
in the system to prevent debris from
entering the heat exchanger.
As is true with all heat exchangers,
it is very important to pipe flat plate
heat exchangers so that the two
liquids pass through it in opposite
directions (e.g., counterflow). In most
cases, each fluid passage through
a flat plate heat exchanger leads
back to two connections, one above
the other, at one end of the heat
exchanger, as shown in Figure 2-19.
Flat plate heat exchangers have
several benefits relative to other heat
exchanger designs. They include:
• A very high ratio of surface area to
internal volume. This allows flat plate
heat exchangers to be significantly
smaller than a shell & tube or shell
& coil heat exchanger of equivalent
heat transfer capacity. It also allows
for a rapid response to temperature
changes in either fluid. Small brazed
plate heat exchangers can achieve
steady state conditions within a
few seconds after two stable fluid
streams (e.g., constant flow rate and
stable entering temperature) begin
flowing through them.
• The metal plates can be thinner
than the tubing used in other types
of heat exchangers. This reduces
the thermal resistance between the
two fluids. Later Sections will show
20
how to account for plate thickness when determining heat
exchanger performance.
• The patterns used on the plates allow for relatively high
turbulence, which increases convective heat transfer and
potentially reduces the required internal surface area of
the heat exchanger. Higher turbulence also decreases the
potential for dirt or other fouling materials to stick to the plates.
FAN-FORCED LIQUID-TO-AIR HEAT EXCHANGERS
One could think of any hydronic heat emitter, be it a cast
iron radiator, heated floor slab or a finned-tube baseboard,
as a liquid-to-air heat exchanger. All of these heat emitters
— to some degree — rely on natural convection heat
transfer to move heat from a stream of water to room air.
These heat emitters have been discussed in several past
issues of idronics. In this issue, the discussion focuses on
devices that use a fan or blower to create airflow through
the heat exchanger.
There are many situations in which the heat in a stream of
water needs to be transferred directly to air within an interior
space. Likewise, nearly all hydronic chilled-water cooling
systems need a means of extracting heat and moisture
from inside air. Many products have been developed for
these applications using forced convection heat transfer on
both the water and air side of the heat exchanger. They can
be categorized as follows:
• Fan-coils
• Air handlers
Figure 2-29
FAN-COILS
A fan-coil is fundamentally a combination of a water-to-air
heat exchanger, known as a “coil,” with a fan or blower
that creates forced convection on the air side of that coil.
In general, fan-coils are designed to heat or cool individual
spaces within a building. Multiple fans coils can be used
to create heating and cooling zones within a building.
All fan-coils can be used for heating when supplied
with heated water. Only fan-coils equipped with internal
condensate drip pans can be used for cooling when
supplied with chilled water.
Figure 2-29 shows the internal components of a modern
“console” fan coil. The coil consists of copper tubing routed
through closely spaced aluminum fins. It’s visible near the
top of the unit. A tangential blower wheel located below
the coil and rotated by a high-efficiency, electronically
commutated motor draws room air into the cabinet from
a few inches above floor level. This air is blown across the
coil and discharged through the upper grill.
Beyond this basic heat transfer and air movement
functionality are controls that vary from one manufacturer
to another. In modern fan-coils, these controls can
regulate blower speed and operate the fan-coil based on
specific setpoint temperatures for heating and cooling
and time of day.
Another form factor that is currently used for fan-coils is
called a “high wall cassette,” an example of which is shown
in Figure 2-30.
High wall cassettes are equipped with condensate drain
pans and can be used for hydronic heating or cooling. They
draw room air into the top of the cabinet and discharge air
through a lower opening equipped with oscillating vanes
to dispense air in different directions. As shown, the unit
in Figure 2-30 is off. The air discharge slot at the bottom
of the unit is covered by a motorized damper.
Figure 2-30
Courtesy of Myson
21
Figure 2-31
supply duct
blower
coil (4-tube pass)
return duct
air filter
drip pan
condensate drain
(with trap)
High wall cassettes are typically turned
on and off using a handheld remote. The
speed of the fan can be adjusted from
the remote or automatically for specific
modes of operation. For example, a low
fan speed is used when the unit is set for
dehumidification mode, allowing the coil to
extract more moisture from the air stream.
Figure 2-32
2-pipe air handler
in heating mode
2-pipe air handler
in cooling mode
Some fan-coils are available for recessed
mounting in wall cavities. Some have
air filters, while others do not. They are
available in a wide range of heating and
cooling capacities depending on size and
source water temperature.
AIR HANDLERS
The combination of a water-to-air heat
exchanger and blower is also the basis
for a category of devices known as air
handlers. Figure 2-31 shows a typical
vertically oriented hydronic air handler.
Air handlers are also available with
horizontally oriented cabinets.
Air handlers, as well as some fan-coils,
are available in either “2-pipe” or “4-pipe”
configurations. These configurations are
shown in Figure 2-32.
heated water
(supply & return)
heated water
(supply & return)
4-pipe air handler
in heating mode
OFF
zone
valves
chilled water
(supply & return)
4-pipe air handler
in cooling mode
OFF
chilled water
(supply & return)
22
Figure 2-33
Notice that the direction of water
flow through the coils in both 2-pipe
and 4-pipe air handlers is opposite
the direction of airflow. This is
important in achieving counterflow
heat exchange, and thus the highest
possible rate of heat transfer for a
given set of operating temperatures
and flow rates.
A “2-pipe” air handler has a single
coil heat exchanger. That coil can
provide heating or cooling depending
on the temperature of the water
stream supplied to it. Two-pipe
air handlers are supplied from a
distribution system that has a single
supply main and single return main.
This distribution system can operate
with heated water or chilled water, but
not both at the same time. Two-pipe
air handlers are used in applications
where the entire distribution system
operates in a single mode (e.g.,
heating mode or cooling mode) at
any given time.
Four-pipe air handlers have two
water-to-air coil heat exchangers.
One coil is piped to a distribution
system for chilled water. The other
coil is piped to a separate distribution
system for heated water. Any 4-pipe
air handler connected to these two
independent distribution systems
can operate in either heating mode
or cooling mode. The operating
mode is usually determined by a
thermostat in the space served by
the air handler. When heating is
needed, a zone valve opens to allow
hot water to flow through the heating
coil in the air handler. The blower is
also turned on. When the thermostat
calls for cooling, a separate zone
valve opens to allow chilled water to
flow through the chilled water coil,
along with blower operation. Fourpipe
distribution systems connected
to multiple 4-pipe air handlers allow
for simultaneous heating and cooling
in different areas of a building.
Figure 2-34
Figure 2-33 shows a partially
installed 4-pipe air handler that’s
suspended from a concrete slab
and above what will eventually
be the room’s ceiling. This is a
common air handler configuration
for buildings such as hotels using a
4-pipe distribution system.
The four copper tubes seen at the
left of the air handler cabinet provide
supply and return water flow to the
heating and cooling coils. They are
fitted with adapters to connect to
PEX distribution tubing. Each coil
circuit is controlled by a Z-one
zone valve within the cabinet. Each
coil circuit is also equipped with a
QuickSetter flow-balancing valve
within the cabinet. The condensate
drip pan, which is connected to a
3/4” PVC drainpipe, is visible at the
base of the unit.
23
One fundamental distinction between
fan coils and air handlers is that the
latter is typically designed to connect
to a ducted air delivery system. This
allows the air stream created by the
blower to be simultaneously delivered
to several locations in a building.
Another distinction is that most air
handlers are intended to be installed
in non-occupied locations such as
mechanical rooms, interior soffits,
basements or attics. They are not
designed with aesthetic details to
make them acceptable in finished
occupied spaces.
Figure 2-34 shows a small,
horizontally oriented, 2-pipe air
handler connected to a ducted
forced-air delivery system.
This unit is mounted in an accessible
area within a conditioned attic. A
removable panel will eventually be
installed to conceal the unit. The coil in
this air handler is supplied by insulated
piping carrying chilled water for
cooling. Because it is installed above
a finished space, this air handler is
mounted over a secondary drain pan
that would capture any condensate
that might leak from the unit’s primary
condensate drip pan. The condensate
formed on the coil during cooling
mode is carried to a drain through the
small PVC pipe seen at the lower right
of the unit.
Small air handlers that are commonly
used in homes or small commercial
buildings may have heating and
cooling capacities ranging from about
12,000 Btu/hr (1 ton) to about 60,000
Btu/hr (5 tons). However, many air
handlers are available with heating
and cooling capacities much higher
than those of fan-coils.
Large air handlers can have heating
and cooling capacities of several
million Btu/hr. These large units are
typically custom-built for specific
applications. They typically have
separate water-to-air coils for heating
and cooling, separate blowers for the
supply and return air streams, energy
recovery devices, and more elaborate
air-filtering systems compared to small
air handlers. Figure 2-35 illustrates
some of these internal details.
The thermal performance of fan-coils
and air handlers will be discussed in
Section 4.
GREYWATER HEAT RECOVERY
HEAT EXCHANGERS
Heat exchangers are also used in
plumbing applications. One example
is recovering heat from domestic
hot water that has passed through a
fixture such as a lavatory or shower.
This water is commonly referred to as
“greywater.”
Much of the heat in domestic hot water
remains in that water as it goes down
the drain. In most buildings, this heat is
simply carried into the sewer. However,
it is possible to recover up to 40% of
this otherwise wasted heat using a
greywater heat exchanger, an example
of which is shown in Figure 2-36.
Figure 2-35
DOAS air handler
rotating desiccant wheel
(exchanges enthalpy
b/w air streams)
exhaust air
heat &
moisture
flow
filter
return air from
building
outdoor air
delivered air
(very low humidity)
filter
pre-conditioned air
heating coil
chilled water coil
supply duct
to building
24
Figure 2-36
Greywater heat exchangers are
simple devices. They are fitted
into a vertical section of greywater
drainage piping. That pipe should be
configured to carry drainage water
from lavatories, showers, tubs and
clothes washers, but not drainage
from garbage disposals, toilets or
bidets.
These heat exchangers consist of
an inner copper pipe that’s wrapped
with one or more coils of tightly fitting,
partially flattened copper tubing that
is bonded to the outer surface of
the inner pipe. The construction of
a typical greywater heat exchanger
is shown in Figure 2-37. A typical
installation configuration is shown in
Figure 2-38.
Greywater heat exchangers leverage
the simultaneous flows of cold water
and hot water to fixtures such as
lavatories and showers. As hot
domestic water is being used, cold
domestic water enters the bottom
connection of the heat exchanger
and passes upward in a counterflow
direction to the greywater. Heat from
the greywater transfers through the
copper tube walls and preheats cold
domestic water. The two streams of
water are always separated by two
Source: Ecoinnovation.ca
Figure 2-37
outer copper helix
(contains domestic water)
Fernco® fittings
(top and bottom)
inner copper pipe
(contains greywater)
partially flattened
copper tubing
domestic water
inner copper pipe
warm greywater
(clings to inner wall)
25
Figure 2-38
(120 ºF)
hot
water
thermostatic
electric
tankless
water heater
(75 ºF)
preheated "cool" water
greywater
heat
recovery
exchanger
shower mixing valve
showerhead
warm
"greywater"
(from fixtures)
(99 ºF)
copper tube walls. A potential leak in
either the inner pipe or outer coil would not
cause contamination of domestic water.
In Figure 2-37, some of the preheated
water leaving the greywater heat exchanger
is piped directly to the “cold” water port
of the shower valve. This reduces the
required flow of fully heated domestic hot
water to the shower valve.
Under typical operating conditions, the
entering cold domestic water will be
warmed 20º to 25ºF before it exits the
heat exchanger. Thus, water entering the
building at 50ºF, would be preheated to a
temperature of 70º to 75ºF before entering
the water heater. This reduces the water
heating load by 29% to 36%, assuming
a final desired delivery temperature of
120ºF. This is a significant contribution
for a simple, unpowered and relatively
inexpensive device.
Greywater heat exchangers should only
be installed in vertical drainage pipes.
They rely on the “film effect” of greywater
passing through a vertical drainage pipe.
Most of this water clings to the pipe wall,
rather than falling through the air space
inside the pipe. This is ideal from the
standpoint of extracting heat from the
inner copper pipe.
cold
water
(50 ºF)
(74 ºF)
cool
"greywater"
(to sewer)
26
3. HEAT TRANSFER FUNDAMENTALS
Two fundamental concepts apply to all situations involving
heat transfer:
1. Heat always moves from a material at some temperature
to another material at a lower temperature.
2. The rate of heat transfer depends on the temperature
difference between the two materials. The greater this
difference, the higher the rate of heat transfer.
To understand the processes at work in different types
of heat exchangers, it’s essential to understand the three
modes of heat transfer. A well-established scientific
principal is that all heat movement occurs by one or more
of the following modes:
• Conduction
• Convection
• Thermal radiation
CONDUCTION
Conduction heat transfer takes place within solid materials.
It’s the result of atomic vibrations within those materials.
The atoms in all materials having temperatures above
absolute zero (-458ºF) vibrate to some extent. The higher
the material’s temperature, the more vigorous the atomic
vibrations.
When heat is added to a solid material, the atomic vibrations
become more energetic. These vibrations spread out
across trillions of atoms that are bonded together, moving
towards material at a lower temperature.
Figure 3-1
The physical property that determines how well a
material transfers heat by conduction is called its thermal
conductivity. The higher a material’s thermal conductivity
value, the faster heat can pass through it, with all other
conditions being equal. The thermal conductivity of a
material is usually determined by testing. The thermal
conductivity of many materials is listed in references such
as the ASHRAE Handbook of Fundamentals.
Heat moving through cast iron that separates the
combustion gases from the water inside a boiler is an
example of conduction heat transfer. Heat passing from
tubing embedded in a concrete slab to the surface of that
slab is another example. Heat passing from a person’s
hand to the cold metal handle of a snow shovel is another.
The rate of heat transfer by conduction is directly
proportional to both the temperature difference across
the material and its thermal conductivity. It is inversely
proportional to the thickness of a material. Thus, if one were
to double the thickness of a material while maintaining the
same temperature difference between its sides, the rate of
heat transfer through the material would be cut in half.
Materials such as copper, aluminum and steel have
relatively high thermal conductivities. Copper and aluminum
in particular are frequently used in applications where high
rates of conduction heat transfer are necessary.
The rate of conduction heat transfer depends on the
thermal conductivity of the material, its thickness, the
area across which heat is passing, and the temperature
difference across it. The relationship between these
quantities is given i29 as formulas Formula 3-1.
Formula 3-1:
⎛
q = A
k ⎞
⎝
⎜
∆ x ⎠
⎟ (∆T )
Where:
q = rate of heat transfer through the material (Btu/hr)
k = thermal conductivity of the material (Btu/°F•hr•ft)
∆x = thickness of the
q = A material R (∆T ) in the direction of heat flow
(ft)
∆T = temperature difference across the material (°F)
A = area heat flows across (ft 2 )
R = ∆ x
The mathematical term (k/∆x) k in Formula 3.1 is sometimes
called a heat transfer coefficient and is represented by the
symbol (U). The reciprocal of the heat transfer coefficient U
(e.g., 1/U) is called “thermal resistance,” and is represented
by (R). q = hA(∆T )
( )
q = sAF T 4 4
− T
12 1 2
⎡ 1 1 ⎤
⎢ + −1 ⎥
27
i29 formulas
⎛
Formula
q
3-1
= A
k ⎞
⎝
⎜can ∆ x ⎠
⎟ (∆T therefore ) be
modified into Formula 3-2:
Formula 3-2:
q = A R (∆T )
Where:
q = rate of heat transfer through the
material (Btu/hr)
R = ∆ x
R = thermal i29 resistance formulas
k (or “R-value”)
of a material (ºF•hr•ft 2 /Btu)
∆T = temperature difference across
the material (°F) ⎛
A = area heat q q = flows AhA(∆T k ⎞
across (ft 2 )
⎝
⎜
∆ x ⎠
⎟ ) (∆T )
The R-value of a material can be
determined based on its thermal
conductivity (k), and its thickness
(∆X). The relationship q = is given as
Formula 3-3. q = sAF A T 4 4
− T
R (∆T 12 ( )
1 2 )
⎡ 1
+ 1 ⎤
⎢ −1⎥
Formula 3-3: ⎣ e 1
e 2 ⎦
R = ∆ x
k
Where:
Re# = vdD
k = thermal conductivity µ of the
material (Btu/°F•hr•ft)
q = hA(∆T )
∆x = thickness of the material in the
direction of heat flow (ft) ⎛ 1 ft ⎞
d = (0.811 in)
The U-values and R-values
⎝
⎜
12 in⎠
⎟
of
materials used to build heat
exchangers q are = sAF critically T 4 4
12 ( − T
1 important.
2 )
In general, materials ⎡ 1and thicknesses
that provide high ⎛ U-value, and thus
low R-value, v allow = 0.408 + 1 ⎤
⎢ ⎞
−1
⎝
⎜ for higher thermal
d 2 ⎠
⎟ f
⎥
⎣ e 1
e 2 ⎦
performance when heat needs
to pass through a solid material
separating the two fluids.
Re# = vdD
CONVECTION
⎛
v = 0.408 µ ⎞
⎛
Convection heat ⎝
⎜
transfer d 2 ⎠
⎟ f occurs = 0.408
⎜
⎝
⎜ 0.811 as
the result of fluid movement. The fluid
⎛ 1 ft ⎞
can be a liquid d or = a (0.811 gas. in)
⎝
⎜
12 in⎠
⎟
When heated, fluids expand. This
lowers their density relative to
surrounding cooler fluid. Lowered
density increases ⎛ buoyancy, which
causes the v warmer = 0.408 ⎞
fluid to rise.
Examples of the
⎝
⎜
latter
d 2
include
⎠
⎟ f
warm
air rising toward the ceiling in a room
and heated water rising to the upper
portion of tank-type water heater.
Both processes occur without
circulators or blowers. As such, they
are examples of “natural” convection.
Convection heat transfer is also
responsible for moving heat between
a fluid and a solid.
For example, consider water at
100ºF flowing along a solid surface
that has a temperature of 120ºF. The
cooler water molecules contacting
the warmer surface absorb heat from
that surface. These molecules are
churned about as the water moves
along. Molecules that have absorbed
heat from the surface are constantly
being swept away from that surface
into the bulk of the water stream and
replaced by cooler molecules. One
can envision this form of convective
heat transfer as heat being “scrubbed
off” the surface by the flowing water.
The speed of the fluid moving over
the surface greatly affects the rate of
convective heat transfer.
Most people have experienced “wind
= 0.06758 chill” ftas cold outside air blows past
( ) 2
⎞
⎟
⎠
⎟
Figure 3-2
5 = 3.1
ft
sec
= 0.06758 ft
them. They feel “colder” compared
to how they would feel if standing
in still air at the same temperature.
The faster the air blows past them,
the greater the rate of convective
heat transfer between their skin or
clothing surfaces and the air stream.
Although the person may feel as if the
moving air is colder, it isn’t. Instead,
they’re experiencing an increased
rate of heat loss due to enhanced
convective heat transfer. Achieving
the same cooling sensation from
calm air would require a much lower
air temperature.
Convective heat transfer increases
with increasing fluid speed. This
happens because a layer of fluid
called the “boundary layer,” which
clings to surfaces, gets thinner as
the fluid’s velocity increases (see
Figure 3-2).
The thinner the boundary layer,
the lower the thermal resistance
between the core of the fluid stream
and the surface. Less thermal
resistance allows for higher rates
of heat transfer between the fluid
molecules in the core of the stream
and the tubing wall.
tube wall
boundary layer of slow moving fluid limits convection
velocity profile of fluid
Note low velocity near tube wall.
"core" of flow stream
heat flows from water,
through boundary layer, tube
wall, into surrounding material
28
⎛
v = 0.408 ⎞
⎛
⎝
⎜
d 2 ⎠
⎟ f = 0.408
⎞
ft
⎜ ⎟ 5 = 3.1
⎝
⎜ 0.811 ⎠
⎟ sec
( ) 2
Figure 3-3
Heat output (Btu/hr/ft)
600
500
400
300
200
100
flow rate = 4 gpm
flow rate = 1 gpm
finned-tube baseboard
0
110 120 130 140 150 160 170 180
Average water temperature (ºF)
exchanger, the higher the rate of heat
transfer between those fluids. From
the standpoint of heat transfer only,
there is no such thing as the water
moving “too fast” through any type
of heat exchanger. However, other
considerations, such as pumping
power, noise or long-term erosion
of materials, impose practical
limits on flow velocity through heat
exchangers.
One example of increasing convective
heat transfer with increasing flow rate
can be found in heat output ratings
for the liquid-to-air heat exchanger in
a small fan-coil unit. Figure 3-4 shows
how heat output increases with
increasing water flow rate through
the coil heat exchanger, while the
inlet water temperature, the incoming
air temperature and the airflow rate all
remain constant.
The heat output rate increases
rapidly at low flow rates. It continues
to increase as flow rate increases, but
the rate of increase is much slower
The relationship between convective
heat transfer and fluid speed across
a surface can be seen in many
hydronic systems. For example, the
faster water flows through a finnedtube
baseboard, the higher it’s heat
output, assuming all other conditions
remain the same. This effect is evident
in the heat output rating data listed
by manufacturers of finned-tube
baseboard. Figure 3-3 shows a plot
of this data for a typical residential
finned-tube baseboard.
At any given water temperature, the
heat output occurring at a flow rate of
4 gallons per minute (gpm) is slightly
higher than at a flow rate of 1 gpm.
This increase is due to a thinner
boundary layer between the bulk of
the fluid stream and the inner tube
wall of the finned-tube element.
In general, the faster one or both
fluids flow through any type of heat
Figure 3-4
Heat output (Btu/hr)
4500
4000
3500
3000
2500
2000
1500
1000
500
0
0 0.5 1 1.5 2 2.5 3 3.5 4 4.5 5
water flow rate (gpm)
29
Figure 3-5
Figure 3-6
internal coil heat exchanger
circulator
heat
source
indirect
water
heater
Forced convection
(inside coil surface)
Natural convection
(outside coil surface)
Forced convection
(inside tube surface)
Natural convection
(outsside tube surface and fin surface)
warm air out
hot
water
in
cool air in
finned-tube baseboard
warm
water
out
aluminum
fins
at higher flows. This demonstrates that there are practical
limits on how much the heat output of a hydronic heat
emitter can be increased based on higher flow rates.
NATURAL vs. FORCED CONVECTION
When fluid motion is caused by a circulator, a blower, a fan
or any other powered device, the convective heat transfer
is called “forced convection.” When the fluid motion is
strictly the result of buoyancy differences within the fluid,
the convective heat transfer is called “natural convection.”
There are many types of heat exchangers that operate with
natural convection on one surface and forced convection
on the other side of that surface. One example is at the
external surface of an internal coil heat exchanger within
an indirect water heater. Another is the external surface of
finned-tube baseboard, as shown in Figure 3-5.
Natural convection is typically a much “weaker” form of
heat transfer compared to forced convection. This is the
result of much slower fluid motion created by buoyancy
differences in the fluid versus faster fluid motion created
by a circulator or blower. ⎛ Slower fluid motion increases
boundary layer thickness, q = A
k ⎞
which creates greater resistance
to heat transfer between ⎝
⎜
∆the x ⎠
⎟ (∆T )
bulk of the fluid and the
surface.
The difference between forced convection and natural
convection explains why a small wall-mounted fan-coil
that’s only 18 inches wide can provide the equivalent
q = A heat output of 10+ feet
R
of (∆T finned-tube ) baseboard when
both are operating at the same water supply temperature
and flow rate. The rate of heat transfer from the finnedtube
element in the baseboard is limited by natural
convection heat transfer between its outer surfaces and
the surrounding air.
The rate of convective heat transfer can be estimated using
Formula 3-4.
Formula 3-4:
i29 formulas
R = ∆ x
k
q = hA(∆T )
Where:
q = rate of heat transfer by convection (Btu/hr)
h = convection coefficient (Btu/hr•ft 2 •ºF)
A = area over which fluid contacts a surface with which it
exchanges heat (ft 2 )
∆T = temperature difference q = sAF T 4 4
12 ( − T
1
between bulk 2 )
fluid stream and
surface (ºF)
⎡ 1
+ 1 ⎤
⎢ −1⎥
⎣ e 1
e 2 ⎦
Although Formula 3-4 is relatively simple, determining the
value of the convection coefficient (h) is often a complex
30
Re# = vdD
µ
process. The value of (h) varies widely depending on
the “geometry” of the surface relative to the fluid, the
speed of the fluid and the physical properties of the
fluid (e.g., thermal conductivity, specific heat, density
and viscosity). In many cases, the value of (h) needs to
be determined experimentally. Heat transfer textbooks
sometimes list values of (h) or methods for determining
(h) for very specific and simplified situations. The table in
Figure 3-6 shows how wide the range of (h) values can
be for specific situations.
THERMAL RADIATION
Heat transfer by thermal radiation is less intuitive compared
to conduction and convection. To many, the word “radiation”
denotes an undesirable or harmful effect associated with
nuclear radiation. However thermal radiation and nuclear
radiation are very different. Human skin as well as clothing
surfaces emit thermal radiation to any surrounding surface
that’s at a lower temperature. A lightly clothed human
body releases over half of its metabolic heat production as
thermal radiation.
It’s helpful to think of thermal radiation as light. More
specifically infrared light. Human eyes cannot see infrared
light. However, just like visible light, infrared light (a.k.a.
thermal radiation) travels outward from its source in straight
lines at the speed of light (186,000 miles per second).
Although it cannot “bend” around corners, thermal
radiation can be reflected by some surfaces. It also travels
equally well in any direction, from some surface that emits
it to another surface that absorbs it. For example, just as a
ceiling-mounted lighting fixture can shine visible light onto
objects below, thermal radiation can shine down from a
heated ceiling and be absorbed by objects below.
The instant thermal radiation is absorbed by the surface of
an object, it becomes heat that warms that object.
Unlike conduction or convection, thermal radiation
needs no material (e.g., a fluid or solid) to transfer heat
from one location to another. Thermal radiation cannot
pass through a solid (as radiation). It can, to differing
extents, pass through gases with minimal warming effect
on those gases. Intense sunlight passing through very
cold air in the upper layers of the earth’s atmosphere is
an example of the latter.
Consider a person kneeling a few feet away from a campfire
on a cold winter day. If pointed toward the fire, their face
likely feels warm, even though the air surrounding them
is cold. This sensation is the result of thermal radiation
emitted by the fire traveling through the cold air and being
absorbed by their exposed skin. The air between the fire
and the person absorbs very little of the energy being
Figure 3-7
transferred from the flames to the person’s face. Likewise,
thermal radiation emitted from the warm surface of a heat
emitter such as a panel radiator can pass through the air
in a room without first heating that air. When the thermal
radiation strikes another surface in the room, most of it is
absorbed. At that instant, the energy carried by the thermal
radiation becomes heat.
Every surface continually emits thermal radiation to any
cooler surface within sight of it. The surface of a heat
emitter that is warmer than our skin or clothing surfaces
transfers heat to us by thermal radiation. Likewise, our
skin and clothing continually give off thermal radiation to
any surrounding surfaces at lower temperatures. A lightly
clothed person standing next to a large cold window
surface emits significant thermal radiation to that cold
surface. This eventually leads to discomfort, even when
the air temperature surrounding the person is in the normal
comfort range of 68-72ºF.
When thermal radiation strikes an opaque surface, part
of it is absorbed as heat and part is reflected away from
the surface. The percentage of incoming radiation that
is absorbed or reflected is determined by the optical
characteristics of the surface and the wavelength of the
radiation. Most interior building surfaces absorb the majority
of thermal radiation that strikes them. The small percentage
that is reflected typically strikes another surface within the
room where most of it will be absorbed, and so on. Very
little, if any, thermal radiation emitted by warm surfaces in a
room escapes from the room.
The rate at which thermal radiation transfers heat between
two surfaces depends upon their temperatures, an optical
property of each surface called emissivity, as well as the
angle and distance between the surfaces.
31
R = ∆ x
k
The rate of heat exchange by thermal radiation between
two parallel flat q = surfaces hA(∆T having ) the same area can be
estimated using Formula 3-5.
Formula 3-5:
( )
q = sAF T 4 4
− T
12 1 2
⎡ 1
+ 1 ⎤
⎢ −1⎥
⎣ e 1
e 2 ⎦
Where:
q = rate of heat transfer from hotter to cooler surface by
thermal radiation Re# (Btu/hr) = vdD
s = Stefan Boltzmann constant
µ
= 0.1714x10 -8 Btu/
hr•ft 2 •ºR 4
F 12 = shape factor between the two surfaces (unitless)
A = area of either surface (ft 2 )
⎛ 1 ft ⎞
T 1 = absolute temperature d = (0.811 of in) the hotter surface (ºR)
T 2 = absolute temperature of the ⎝
⎜
cooler 12 in⎠
⎟ = 0.06758 ft
surface (ºR)
e 1 = emissivity of the hotter surface (unitless)
e 2 = emissivity of the cooler surface (unitless)
It’s also important to understand that the temperatures (T1)
and (T2) in Formula 3-5 must be absolute temperatures.
Temperatures in ºF can be converted to absolute
temperatures in degrees Rankine (ºR) by adding 458
degrees. Thus, 32ºF becomes 32 + 458 = 490ºR.
The mathematical result of the calculation (T 1 4 -T 2 4 )
changes much more than the simple ∆T term used in the
formulas for conduction and convection. For example,
consider two surfaces exchanging radiant heat with
temperatures of 100ºF and 80ºF. These temperatures
would convert to 558ºR and 538ºR. The difference between
these temperatures would be only 20ºR, the same as the
difference between 100ºF and 80ºF. However, when these
Rankine temperatures are used in Formula 3-5 the resulting
number for the term (T 1 4 -T 2 4 ) is 131,700,000ºR 4 .
Due to mathematical complexities, as well as variability
or uncertainty in properties such as surface emissivities,
theoretical calculations of radiant heat transfer are often
limited to relatively simple situations.
This formula is more complex that those used for estimating
conduction and convection heat transfer.
⎛
v = 0.408 ⎞
⎝
⎜
⎠
⎟ f
d 2
The value of the “shape factor” (F 12 ) is a number between
0 and 1.0. It’s determined based on the relative angle and
distance between the two surfaces exchanging radiant
heat. Heat transfer textbooks give specific methods for
finding values of the shape factor (F 12 ) for different surfaces
⎛
and orientations. v = For 0.408 ⎞
⎛
two parallel planes having infinite
width and depth, the ⎝
⎜
value of ⎠
⎟ f = 0.408
⎞
⎜ ⎟
the shape factor (F 12 ) is 1.0.
d 2
ft
5 = 3.1
⎝
⎜ ( 0.811) 2
⎠
⎟ sec
Factors e 1 and e 2 are the emissivities of surfaces 1 and 2.
Emissivity is a surface property determined experimentally
based on how well the surface emits thermal radiation.
It must be a number between 0 and 1. A high value
indicates that the surface is a good emitter, and vice versa.
Emissivity values for various surfaces can be found in
references such as heat transfer handbooks. Interestingly,
the emissivity of a surface is not necessarily correlated with
its color. A rough metal surface coated with white enamel
paint has an emissivity of 0.91, and a flat black painted
surface has an emissivity of 0.97. Freshly fallen snow can
have an emissivity over 0.90. The emissivity of a polished
copper surface is 0.023, while a heavily oxidized copper
surface has an emissivity of 0.78. Most highly polished
metal surfaces have low emissivities, and thus would not
be good choices for the surface of a hydronic heat emitter
that’s expected to radiate heat into a room.
32
4. THERMAL CHARACTERISTICS OF HEAT EXCHANGERS
Figure 4-1
laminar flow
through a round pipe
small (bouyancy neutral)
particles carried along by fluid
boundary layer
streamlines
boundary layer
The thermal performance of heat
exchangers is very dependent on
the physical properties of the two
fluids passing through them. It’s
also dependent on the “dynamic
condition” of those fluids. The latter
refers to the flow direction, velocity
and degree of turbulence present in
the fluids exchanging heat.
LAMINAR vs. TURBULENT FLOW
It’s helpful to imagine liquids flowing
through pipes or along the surface
of a plate as a stream containing
millions of tiny buoyancy-neutral
“particles.” The path that any one of
these particles takes as it moves is
called a streamline.
In some situations, the streamlines
of all the fluid particles are parallel to
each other, like multiple lanes of traffic
on a major highway. The particles
near the center of the pipe, or those
farther away from a flat surface, move
faster than those near the pipe wall
or the flat surface, but they don’t
“change lanes.” This type of fluid
movement, represented by Figure
4-1, is called laminar flow.
Laminar flow can be desirable or
undesirable depending on where it
occurs in hydronic systems, as well
as the intended purpose of those
laminar flow
over a flat surface
small (bouyancy neutral)
particles carried along by fluid
boundary layer
horizontal plate
systems. If the goal is to move fluid
long distances with minimal pumping
power, laminar flow is desirable
because it reduces pumping power.
However, if the goal is to maximize
heat transfer from a flowing fluid,
laminar flow is very undesirable.
Because the fluid particles move in
parallel streamlines during laminar
flow, there is very little mixing between
them. This increases the thickness of
the boundary layer between the bulk
of the fluid stream and the wall of the
pipe or surface of a plate. Thicker
boundary layers decrease convective
heat transfer.
Figure 4-2
streamlines
turbulent flow
through a round pipe
small (bouyancy neutral)
particles carried along by fluid
boundary layer
streamlines
boundary layer
Turbulent flow causes the streamlines
of the imaginary fluid particles to bend
and twist as the fluid moves down the
pipe or along a surface, as shown in
Figure 4-2.
The erratically shaped streamlines
present during turbulent flow create
good mixing. A small quantity of fluid
that’s close to a solid surface one
instant, can be swept into the bulk
of the fluid stream the next instant.
This decreases the boundary layer
thickness and significantly improves
convective heat transfer.
The convective heat transfer
associated with a relatively viscous
liquid at low temperatures, such as
a 25% solution of propylene glycol
antifreeze at 30ºF, under laminar flow
is only about 7% of the convective
heat transfer of the same fluid
under turbulent flow. This implies
that laminar flow will result in a very
significant drop in heat transfer
between a fluid and the inner wall of a
tube or the surface of a plate. Such a
change could have a major negative
effect on the ability of a hydronic
circuit, such as a ground loop heat
exchanger supplying a geothermal
heat pump, to transfer heat at the
required rate.
turbulent flow
over a flat surface
small (bouyancy neutral)
particles carried along by fluid
streamlines
boundary layer
horizontal plate
33
Although turbulent flow improves convective heat transfer, it
also increases head loss, and thus requires more pumping
power to maintain a given flow rate relative to the power
required for i29 laminar formulas flow. However, the increased heat
transfer capabilities of turbulent flow often far outweigh
the penalty associated with the increased pumping power.
Thus, when possible, all heat exchangers should be
operated with turbulent ⎛ flow of both fluids.
q = A
k ⎞
⎝
⎜
∆ x ⎠
⎟ (∆T )
REYNOLDS NUMBER
It’s possible to predict if flow through a pipe or across a
smooth plate will be laminar or turbulent. It’s based on
calculating a dimensionless quantity called the Reynolds
number of the fluid (abbreviated as Re#).
q = A R (∆T )
The Reynolds number is a ratio of the inertial forces existing
in a flow stream compared to the viscous forces in that
flow stream. It takes several physical characteristics of
the flowing fluid into account, including the fluid’s density
and viscosity, R both = ∆ x
of which are dependent on the fluid’s
temperature. It also
kaccounts for the speed of the fluid and
the geometry of surfaces in contact with the fluid.
When the Reynolds number is low, the inertial forces
are relatively weak q = hA(∆T compared ) to the viscous forces. Any
disturbance to the flow stream that might otherwise induce
turbulence is quickly dampened out by viscous forces, and
thus sustained turbulence cannot exist. However, when the
inertial forces become dominant over the viscous forces,
which would be characterized by higher Reynolds numbers,
the dampening effects of viscosity are not able to prevent
turbulent flow q from = sAF T 4 4
12 ( − T
1 2 )
being established and maintained.
For flow through a round pipe, the Re# is calculated using
Formula 4-1.
⎣ ⎦
Formula 4-1:
⎡ 1
+ 1 ⎤
⎢ −1⎥
e 1
e 2
Re# = vdD
µ
where:
v = average flow velocity of the ⎛fluid 1 ft (ft/sec) ⎞
d = (0.811 in)
d = internal diameter of pipe (ft)
⎝
⎜
12 in⎠
⎟
D = fluid’s density (lb/ft 3 )
µ = fluid’s dynamic viscosity (lb/ft/sec)
= 0.06758 ft
If the Reynolds number of flow through a round pipe is
over 4,000, the flow will be turbulent. If the Re# is below
⎛
2,300, the flow v = through 0.408 ⎞
the pipe will be laminar. If the Re#
is between 2300 and ⎝
⎜
4000, d 2 ⎠
⎟ f
the flow could be either laminar
or turbulent, and it could switch back and forth.
density (lb/ft3)
Figure 4-3
63
62.5
62
61.5
61
60.5
60
59.5
59
58.5
58
50 100 150 200 250
Here’s an example: Determine the Reynolds number of
water at 140°F flowing at 5 gpm through a 3/4-inch type M
copper tube. Is this flow laminar or turbulent?
Solution: To calculate the Re#, the density and dynamic
viscosity of the water must be determined. The density of
water can be found from Figure 4-3. The dynamic viscosity
of water can be found from Figure 4-4.
Figure 4-4
dynamic viscosity (lb/ft/sec)
0.0009
0.0008
0.0007
0.0006
0.0005
0.0004
0.0003
0.0002
0.0001
0
50 100 150 200 250
34
⎛
v = 0.408 ⎞
⎛
⎝
⎜
d 2 ⎠
⎟ f = 0.408
⎜
⎝
⎜ 0.811
( ) 2
⎞
⎟
⎠
⎟
5 = 3.1
ft
sec
( )
q sAF T 4 4
− T
12 1 2
⎡ 1 q =
+ 1 sAF ⎤
T 4 4
q = hA(∆T ) 12 ( − T
1 2 )
⎢ −1
e 1
e
⎡ 1⎥
⎣ 2 ⎦ + 1 ⎤
⎢ −1⎥
At 140ºF, the density of ⎣ ewater 1
e 2 is read ⎦ from Figure 4-3 as
61.35 lb/ft 3 . From Figure 4-4, the dynamic viscosity of
water at 140ºF is 0.00032 lb/ft/sec.
q
The inside Re# = sAF diameter = vdD T 4 4
12 ( − T
1 2 )
⎡ 1 µ
of a 3/4-inch type M copper tube is
0.811 inches. This + Re# 1 ⎤
must = vdD
⎢ −1⎥be converted
µ
to feet to match the
e
stated units for ⎣ 1
e
Formula 2
4-1: ⎦
⎛ 1 ft ⎞
d = (0.811 in)
⎝
⎜
12 in⎠
⎟ = ⎛0.06758 1 ft ⎞ ft
d = (0.811 in)
⎝
⎜
12 in⎠
⎟ = 0.06758 ft
Re# = vdD
The average flow velocity µ corresponding to a flow rate of 5
gpm can be found using Formula 4-2.
FLOW DIRECTION THROUGH HEAT EXCHANGER
The liquid-to-liquid heat exchangers used in hydronic
systems are configured so that the two fluid streams
flow in mostly parallel directions to each other. The two
parallel streams could be flowing in the same direction or
in opposite directions. When the two streams flow in the
same direction, the configuration is called “parallel flow,”
or sometimes “cocurrent” flow. When the two flow streams
flow in opposite directions, the configuration is called
“counterflow.” These two flow configurations, along with
representative temperature changes of both fluid streams,
are shown in Figure 4-5.
Formula 4-2: ⎛
v = 0.408 ⎞
⎝
⎜
d 2 ⎠
⎟
v = f ⎛ 1 ft ⎞
d = (0.811 in) ⎛ 0.408 ⎝
⎜
12 in ⎞ ⎠
⎟ = 0.06758 ft
⎝
⎜
⎠
⎟ f
d 2
Figure 4-5
T hotin
Where:
v = average fluid velocity (ft/sec)
d = exact inside ⎛⎛
diameter of pipe (inches)
f = flow rate through ⎞
⎛
pipe (gpm)
⎝
⎜
2 ⎠
⎟ f = 0.408
⎞
ft
v = 0.408 ⎞
⎛ ⎜ ⎟ 5 = 3.1
v = 0.408
⎝
⎜
⎞
⎛
( 0.811) 2
⎠
⎟ sec
For 3/4-inch Type M copper ⎝
⎜
d 2
tubing ⎠
⎟ f = 0.408
⎞
⎝
⎜
d 2 ⎠
⎟ f
ft
⎜ ⎟ 5 = 3.1
operating ⎝
⎜ ( 0.811) at 2
⎠
⎟a flow rate sec
of 5 gpm, the average flow velocity is:
hot fluid
cold fluid
T hot
out
T cold
out
⎛
v = 0.408 ⎞
⎛
⎝
⎜
d 2 ⎠
⎟ f = 0.408
⎜
⎝
⎜ 0.811
( ) 2
⎞
ft
⎟ 5 = 3.1
⎠
⎟ sec
The Reynolds number can now be calculated using
Formula 4-1.
⎛
3.1 ft ⎞
⎛
⎝
⎜
sec ⎠
⎟ 0.06758 ft
⎝
⎜
Re# =
lb
0.00032
ft ⋅sec
( ) 61.35 lb
⎞
ft 3 ⎠
⎟
= 40,165
T cold
in
T hotin
T cold
in
T hotin
parallel flow
T hot
out
T cold
out
This value is well above the 4,000 threshold and therefore
the flow is turbulent.
(
Notice LMTD that the = units ∆T ) 1
− ( ∆T )
on the quantities 2
used in Formula 4-1
cancel out completely. ⎡
ln (∆T A valid ) ⎤
1 Reynold number must always
⎢ ⎥
be a number — with ⎣(∆T no ) units, 2 ⎦ regardless of the use of the
IP or SI unit system. When calculating a Reynolds number,
it is good practice to enter the units on all quantities and be
sure they all cancel out.
(
From a
LMTD
practical
= ∆T ) 1
− ( ∆T ) 2
45− 36 9
standpoint, most flows in hydronic heating
and cooling systems ln (∆T will ) =
⎡ ⎤
1
⎢ have ⎥ Re# ln well 45 =
⎡ ⎤ 0.223 = 40.3º F
⎢ above 4,000, and
thus be turbulent. The ⎣(∆T higher
) 2 ⎦
36
⎥
the Re#, ⎣ the ⎦ more intense the
turbulence. Flows through any device that adds or removes
heat from a fluid (e.g., any heat exchanger, hydronic heat
emitter or cooling coil) should always be turbulent.
( ) 1
− ( ∆T ) 2
LMTD = ∆T
⎡
ln (∆T ) ⎤
1
⎢ ⎥
⎣(∆T ) 2 ⎦
36 − 45
=
ln 36 =
⎡ ⎤
⎢
⎣ 45
⎥
⎦
−9
−0.223 = 40.3º F
T cold
out
T hotin
T cold
out
hot fluid
cold fluid
counterflow
T hot
out
Tcold
in
T hot
out
Tcold
in
35
One of the fundamental characteristics of heat exchangers
is that the rate of heat transfer depends on the temperature
difference between the fluid providing the heat and the
fluid absorbing that heat. The greater this temperature
difference, the higher the rate of heat transfer.
Figure 4-6
122 ºF
77 ºF
It’s apparent that the temperature difference between the
hot fluid and cool fluid in the heat exchangers represented
in Figure 4-5 varies considerably depending on the location
within the heat exchanger where this difference is measured.
For the parallel flow heat exchanger, the temperature
difference at the left side, where both fluids enter, is very
large. But this temperature difference decreases rapidly
as the fluids exchange heat and move toward the outlet
ports. There’s also a variation in the temperature difference
⎛
3.1 ft ⎞
⎛
between the fluids as they move through the counterflow
⎝
⎜
sec ⎠
⎟ ( 0.06758 ft) 61.35 lb ⎞
⎝
⎜
ft 3 ⎠
⎟
Re# =
= 40,165
heat exchanger.
lb
⎛ 0.00032
To evaluate the rate of heat transfer across the entire
3.1 ft ⎞
⋅sec
⎛
⎝
⎜
sec ⎠
⎟ ( 0.06758 ft) 61.35 lb ⎞
⎝
⎜
ft
heat exchanger, it is necessary to establish an “average”
104 ºF 68 3 ⎠
⎟
Re# =
ºF = 40,165
lb
temperature difference. This average temperature
0.00032
difference could be defined as a value that if present at all
locations that separate the fluids within the heat exchanger To calculate LMTD = the ∆T ft ⋅sec
(
LMTD,
) 1
− ( ∆T
assume
) 2 that (∆T) 1 is for the top end
would result in the same overall heat exchange rate as of the heat exchanger. ⎡ Thus, (∆T) 1 = 122-77 = 45 ºF. This
will occur given the ⎛actual and highly variable temperature means that (∆T) 2 is at bottom end of the heat exchanger.
differences. 3.1 ft ⎞
Thus, (∆T) 2 = 104-68 = 36ºF. Putting these values for (∆T)
⎝
⎜
sec ⎠
⎟ ( 0.06758 ft) 61.35 lb
ln (∆T ) ⎤
1
⎢ ⎥
⎛ ⎞
(∆T )
⎝
⎜
ft 3 ⎠
⎟
(
⎣ 2 ⎦
LMTD = ∆T ) 1
− ( ∆T ) 2
1
Re# =
= 40,165 and (∆T) 2 into Formula ⎡ 4-3 yields:
Heat transfer theory can be used to derive lb this “average”
ln (∆T ) ⎤
1
0.00032
⎢ ⎥
∆T. It’s called the log mean temperature ft ⋅sec difference (LMTD)
(
and can be calculated using Formula 4-3.
LMTD = ∆T ⎣(∆T )
) 1
− ( 2
∆T ⎦
) 2
45− 36 9
ln (∆T ) =
⎡ ⎤
1
⎢ ⎥ ln 45 =
⎡ ⎤ 0.223 = 40.3º F
⎢
Formula 4-3:
LMTD = ∆T
⎣(∆T ) 2 ⎦
36
⎥
⎣ ⎦
( ) 1
− ( ∆T ) (
2
LMTD = ∆T ) 1
− ( ∆T ) 2
45− 36 9
⎡
ln (∆T ) ⎤
1
ln (∆T ) =
⎡ ⎤
1
Now suppose ⎢ that the ⎥ ln 45 =
⎡ ⎤ 0.223 = 40.3º F
ends ⎢
⎢ ⎥
of the heat exchanger
(∆T ) ⎣(∆T ) 2 ⎦
36
⎥
⎣ ⎦
⎣ 2 ⎦
representing (∆T) 1 and (∆T) 2 were reversed. Thus (∆T) 1 =
Where:
36ºF, and (∆T)
(
LMTD = ∆T
2 =
) 1
−
45ºF.
( ∆T
Putting
) 2
36these − 45 values −9 into Formula
LMTD = log mean temperature difference (ºF)
4-3 yields:
ln (∆T ) =
⎡ ⎤
1
(∆T) 1 = temperature difference between the two fluids at
one end of the LMTD heat exchanger
(
= ∆T
⎢ ⎥ ln 36 =
⎡ ⎤ −0.223 = 40.3º F
⎢
) 1
− ∆T
(ºF)
( ) 2
45− 36 9
(∆T) 2 = temperature difference ⎡ between the two fluids at
the other end of the heat exchanger
ln (∆T ) =
⎤ ⎡
1
⎢ ⎥ ln 45 =
⎤ 0.223 = 40.3º F
⎣(∆T ) 2 ⎦
45
⎥
⎣ ⎦
(
LMTD = ∆T ) 1
− ( ∆T ) 2
36 − 45 −9
(∆T ) (ºF) ⎢
⎣ 2 ⎦ ⎣36
⎥
ln (∆T ) =
⎡ ⎤
1
⎦
⎢ ⎥ ln 36 =
⎡ ⎤ −0.223 = 40.3º F
⎢
ln [ ] = the natural logarithm of the quantity in the square
⎣(∆T ) 2 ⎦
45
⎥
⎣ ⎦
brackets.
q ∝ A(LMTD)
Heat transfer theory can also be used to prove a very
Formula 4-3 applies to both parallel flow and counter flow important concept in the application of heat exchangers.
heat exchangers. It’s also ( important to understand that it Namely that the LMTD of a heat exchanger configured for
makes no difference LMTD which = ∆T ) 1
− ( ∆T ) 2
36 − 45 −9
end of the heat exchanger is counterflow will always be higher than that of the same
⎡
assigned to be (∆T) 1 as long ln (∆T as the ) =
⎤ ⎡
1
⎢ ⎥other lnend ⎢
36 =
⎤ −0.223 = 40.3º q F
q ∝= UA(LMTD)
is assigned heat exchanger configured for parallel flow and having the
to be (∆T) 2 .
⎣(∆T ) 2 ⎦ ⎣ 45
⎥
⎦
same q = UA(LMTD) entering conditions for both flow streams.
q = UA(LMTD) = A(LMTD) A(LMTD)
=
Here’s an example: Calculate the LMTD for the heat The higher the LMTD, the R higher the rate of heat transfer.
T
⎡ 1
exchanger shown in Figure 4-6.
This implies that to attain the highest ⎢ possible + R rate of heat
h ffh
+ ∆ x
q ∝ A(LMTD)
transfer, heat exchangers should always h
k + R + 1 ⎤
ffc ⎥
⎣
h
be configured c ⎦
q = UA(LMTD) = A(LMTD) A(LMTD)
=
R T
⎡ 1
+ R
h ffh
+ ∆ x
q = UA(LMTD)
h
k + R + 1 ⎤
⎢
ffc ⎥
⎣
h c ⎦
1
U =
⎡ 1
⎢ + R ffh
+ ∆ x + R ffc
+ 1 ⎤
⎥
36
side supplying heat
side absorbing heat
0.00032 ft ⋅sec
( ) 1
− ( ∆T ) 2
LMTD = ∆T
for counterflow rather ⎡than parallel
flow. This concept ln (∆T ) ⎤
1
should ⎢ always ⎥
R T
= 1 + R
h ffh
+ ∆ x
h
k + R ffc
+ 1 Figure 4-7
h
be observed when ⎣(∆T ) c
planning 2 ⎦
T
and documenting the design of
1 T 2 T 3 T 4 T 5 T 6
any hydronic system using heat
1
∆ x
1
R
exchangers.
h ffh
R
h k
ffc
h
⎛ (
c
LMTD = ∆T ) 1
− ( ∆T )
3.1 ft ⎞
2
45− ⎛ 36 9
DETERMINING OVERALL RATE OF
HEAT TRANSFER (LMTD ln METHOD)
(∆T ) =
⎡ ⎤
1
⎢ ⎥ ln 45 =
⎝
⎜
sec ⎠
⎟ ( 0.06758 ft) 61.35 lb ⎞
⎝
⎜
⎡ ⎤ ft0.223 3 ⎠
⎟ = 40.3º F
Re# =
= 40,165
The heat transfer rate through
(∆T ) lb ⎢
⎣ 2 a ⎦
36
⎥
0.00032 heat ⎣ ⎦
exchanger is directly proportional
ft ⋅sec
T 1
to the LMTD. It is also directly
proportional to the surface area that
bulk temperature
separates the two fluids. Thus, the
of hot fluid
heat transfer rate is proportional ( to the
T 2
T 3
LMTD = ∆T ( ) 1
− ( ∆T ) 2
36 − 45 −9
T4
multiplication LMTD of internal = ∆T )
⎡surface area
T
(A) times the log mean ln (∆T 1
− ( ∆T ) ) = 2
⎤ ⎡
1
temperature temperature at inside
5
⎢ ⎥ ln ⎢
36 =
⎤ −0.223 = 40.3º F
⎡
ln (∆T ) ⎤
1
of boundary layer
difference (LMTD). This ⎣(∆T ⎢
can ) ⎥
2 ⎦
45
⎥
be ⎣ ⎦
⎣(∆T ) 2 ⎦
expressed as the proportionality:
temperature at
T 6
metal surface
Formula 4-4:
q ∝ A(LMTD) (
LMTD = ∆T ) 1
− ( ∆T ) 2
45− 36 9
ln (∆T ) =
⎡ ⎤
1
Where:
⎢ ⎥ ln 45 =
⎡ ⎤ 0.223 = 40.3º F
⎢
⎣(∆T ) 2 ⎦
36
⎥
⎣ ⎦
q = rate q of = heat UA(LMTD) transfer across the
heat exchanger (Btu/hr)
hot fluid
A = internal surface area separating
boundary layer of fluid
the two fluids within the heat
exchanger (ft 2 )
layer of fouling material
(
LMTD = q = log UA(LMTD) mean temperature = A(LMTD) A(LMTD)
=
(undesirable)
= ∆T ) 1
− ( ∆T ) 2
36 − 45 −9
difference established by the current R T 1
+ R
conditions (ºF)
knowing h ffh
+ ∆ x
h
the k convection + R + 1 ⎤
ln (∆T ) =
⎡ ⎤
1
⎢ ⎥ ln 36 =
⎡ ⎤ −0.223 = 40.3º F
⎢
ffc ⎥
hcoefficient
⎣(∆T ) 2 ⎦
45
⎥
⎣ ⎦
c ⎦
for both fluid-to-surface interfaces,
as well as the thermal conductivity of
the wall separating the fluids. In some
cases, additional thermal resistance
effects called “fouling factors” are
This proportionality can be changed
to a formula (e.g., an equality with an
= sign, rather than a proportionality)
by introducing a constant called U.
U q=
∝ A(LMTD)
1
⎡ 1
Formula 4-5: ⎢ + R
h ffh
+ ∆ x
h
k + R + 1 ffc
⎣
h c
q = UA(LMTD)
The constant U is called the overall
heat transfer coefficient of the heat
exchanger. To make the units in
q = UA(LMTD) = A(LMTD)
⎤also included in the mathematical
⎥definition of the U-value.
⎦
Figure 4-7 shows how the U-value
for a heat exchanger where the two
fluids are separated by a flat plate
can be viewed as a sum of several
thermal resistances. A(LMTD)
Formula 4-5 consistent, U must have =
R
units of Btu/hr/ft 2 /ºF. The value of U is T
⎡ 1
The orange + R lines in Figure 4-7
the rate of heat transfer through the represent h ffh
+ ∆ x
h relative k + R + 1 ⎤
⎢
ffc ⎥
⎣
temperature h c ⎦
heat exchanger (in Btu/hr), per square changes. The bulk temperature of
foot of internal surface area, per ºF of the hot fluid is represented by the
log mean temperature difference. The black dot labelled (T1). There is a
U-value of a heat exchanger involves temperature drop from (T1) to (T2)
1
U =
⎡ 1
+ R
h ffh
+ ∆ x
h
k + R + 1 ⎤
⎢
ffc ⎥
⎣
h c ⎦
metal
wall
of HX
temperature at
metal surface
temperature at
inside boundary layer
heat
flow
bulk temperature
of cool fluid
cool fluid
boundary layer of fluid
layer of fouling material
(undesirable)
across the convective boundary layer.
The extent of this temperature drop
depends on the characteristics of the
boundary layer. Laminar flow would
result in a relatively thick boundary
layer and a larger temperature drop
from (T1) to (T2). Highly turbulent
flow would decrease the thickness
of this boundary layer and reduce the
temperature drop across it.
Another temperature drop, (T2) to
(T3), occurs across the resistance
associated with a fouling layer. If the
heat exchanger is new, or completely
clean of any dirt or accumulated films,
the thermal resistance of the fouling
film is zero, and thus the temperature
drop from (T2) to (T3) is also zero. If
operating conditions allow a fouling
film to develop, the thermal resistance
37
⎛
3.1 ft ⎞
⎛
⎝
⎜
sec ⎠
⎟ ( 0.06758 ft) 61.35 lb ⎞
⎝
⎜
ft 3 ⎠
⎟
Re# =
= 40,165
lb
called “fouling factor” 0.00032 increases, and so does the temperature
drop from (T2) to (T3). ft ⋅sec
Another temperature drop occurs across the metal wall
of the heat exchanger that separates the two fluids.
Because of ( the high thermal conductivity of the wall, the
LMTD = ∆T ) 1
− ( ∆T ) 2
temperature drop ⎡ from (T3) to (T4) will be quite small,
perhaps only ln a fraction (∆T ) ⎤
1
⎢ ⎥ of one degree F. In some situations,
this temperature ⎣(∆T )
drop 2 ⎦is so small relative to the resistance
of the fluid boundary layers and fouling films that it’s
deemed insignificant and not factored into the calculation
of the overall heat transfer coefficient (U).
(
LMTD = ∆T ) 1
− ( ∆T ) 2
45− 36 9
Similar temperature drops occur on the right side of the
ln (∆T ) =
⎡ ⎤
1
metal wall, across ⎢ any ⎥ ln 45 =
⎡ ⎤ 0.223 = 40.3º F
(∆T ) fouling ⎢ film and the convective
⎣ 2 ⎦ ⎣36
⎥
⎦
boundary layer of the cool fluid.
Each of the temperature drops occurs because of a
resistance to heat flow. The diagram at the top of Figure
4-7 represents ( these five thermal resistances, with the
LMTD = ∆T ) 1
− ( ∆T ) 2
36 − 45 −9
temperatures (T1 through T6) present between these
resistances. ln The (∆T numerical ) =
⎡ ⎤
1
⎢ ⎥ value ln 36 =
⎡ ⎤ −0.223 = 40.3º F
⎢ of each thermal resistance
can be estimated ⎣(∆T )
based 2 ⎦
45
⎥
on several
⎣ ⎦
factors, such as fluid
properties, flow velocity, temperature, Reynolds numbers
and more. Some of these calculations can be complex and
require iteration.
q ∝ A(LMTD)
The overall rate of heat transfer depends on these five
thermal resistances. Formula 4-5 can be expanded to
show q = UA(LMTD) how these resistances influence the overall rate of
heat transfer. That expansion is shown as Formula 4-6.
Formula 4-6:
q = UA(LMTD) = A(LMTD) A(LMTD)
=
R T
⎡ 1
+ R
h ffh
+ ∆ x
h
k + R + 1 ⎤
⎢
ffc ⎥
⎣
h c ⎦
Where:
q = rate of heat transfer across the heat exchanger (Btu/hr)
A = internal surface area separating the two fluids within
the heat exchanger 1 (ft 2 )
U =
LMTD ⎡ = 1 log mean temperature difference established by
the current + Rconditions (ºF)
h ffh
+ ∆ x
h h = convection h
k + R + 1 ⎤
⎢
ffc ⎥
⎣
h
coefficient on c ⎦ the hot side of the heat
exchanger (Btu/hr/ft 2 /ºF)
R ffh = thermal resistance of fouling factor on hot side
(ºF•hr•ft 2 /Btu)
∆X = thickness of metal wall of heat exchanger (ft)
k = thermal conductivity of metal wall of heat exchanger
(Btu/hr/ft/ºF)
R ffc = thermal resistance of fouling factor on cool side
(ºF•hr•ft 2 /Btu)
h c = convection coefficient on the cool side of the heat
exchanger (Btu/hr/ft 2 /ºF)
q = UA(LMTD) = A(LMTD)
R T
=
Formula 4-7:
1
U =
⎡ 1
+ R
h ffh
+ ∆ x
h
k + R + 1 ⎤
⎢
ffc ⎥
⎣
h c ⎦
A(LMTD)
⎡ 1
+ R
h ffh
+ ∆ x
h
k + R + 1 ⎤
⎢
ffc ⎥
⎣
h c ⎦
From Formula 4-5 it follows that the overall heat transfer
coefficient (U) can be written as Formula 4-7.
The values of U are highly dependent on the fluids
exchanging heat, the flow conditions and the characteristics
of the material of which the heat exchanger is made. For
water-to-water heat exchangers a guideline range for
values of U is 150-300 Btu/hr/ft 2 /ºF.
Numerical values for fouling factors are given in heat transfer
reference books or on websites. The units for fouling
factors are the same as for insulation R-values or other
thermal resistances. In the IP units system, those units
are (ºF•hr•ft 2 /Btu). Figure 4-8 lists some fouling factors
relevant to hydronic heating and cooling applications and
their associated fluids.
Figure 4-8
These values show a range of thermal resistances
associated with different water conditions. Notice the
fouling factor associated with demineralized water is only
about 17 percent of that associated with untreated river
water or untreated cooling tower water. The presence of
minerals in flow streams passing through heat exchangers,
especially when those fluids are at elevated temperatures,
can significantly increase fouling factors and thus decrease
rates of heat transfer. This again points to the importance of
using high-efficiency dirt separators as well as demineralized
water for optimal hydronic system performance.
The following example illustrates how Formula 4-7 can
be used.
A flat plate heat exchanger is used to heat domestic water.
Boiler water enters the primary side of the heat exchanger
at 150ºF and flow rate of 10 gpm. Cold domestic water
38
Figure 4-9
enters the secondary side of the heat exchanger at 50ºF and
6 gpm. The plates in the heat exchanger are 0.02 inches
thick and made of 316 stainless steel having a thermal
conductivity of 29 Btu/hr/ft/ºF. The heat exchanger is new,
and therefore there are no fouling films on either side of the
plates. The heat exchanger is configured for counterflow,
as shown in Figure 4-9. Prior calculations by the designer
established that the convection coefficient on the hot side
of the heat exchanger is 250 Btu/hr/ft 2 /ºF, and on the cool
side is 150 Btu/hr/ft 2 /ºF. Assume that the heat exchanger
is operating under steady state conditions and is insulated
so that external heat loss is negligible. Determine the rate
of energy flow through the heat exchanger and the required
internal area of the plates.
The rate of heat transfer into the heat exchanger can be
determined based on the flow rate, temperature drop
and physical properties of water flowing through the left
side of the heat exchanger at an average temperature of
(150+135)/2 = 142.5 ºF.
The density of water at 142.5ºF is 61.3 lb/ft 3 . The specific
heat of water at this temperature is 1.00 Btu/lb/ºF.
The sensible heat rate equation (Formula 4-8) can be used
to calculate the rate of heat transfer.
Formula 4-8:
150 ºF
side supplying heat
10 gpm
6 gpm
Tout
side absorbing heat
Where:
q = rate of heat transfer (Btu/hr)
D = density of the fluid (lb/ft 3 )
c = Specific heat of the fluid (Btu/lb/ºF)
f = flow rate (gpm)
∆T = temperature change of the fluid (ºF)
8.01 = constant to balance the units on each side of formula
q = (8.01Dc) f (∆T )
Putting the values into Formula 4-8 yields:
q = (8.01Dc) f (∆T ) = (8.01× 61.3×1.00)(10)(150 −135) = 73,650 Btu
q = (8.01Dc) f (∆T )
hr
q = (8.01Dc) f (∆T )
Since the heat exchanger is operating under steady state
conditions, 73650
T and with insignificant external heat loss, the
out
=
+ 50 = 74.6º F
right (8.01× side 62.4 must ×1.00)(6)
q
q = (8.01Dc) be releasing f (∆T
(∆T heat )
) = (8.01× at the same 61.3×1.00)(10)(150 rate as heat −135) =
is entering the left side. This allows the sensible heat rate
q = (8.01Dc) f (∆T ) = (8.01× 61.3×1.00)(10)(150 −135) = 73,650 Btu
equation q = (8.01Dc) to be f used (∆T ) to determine the outlet temperature on hr
135 ºF 50 ºF
the secondary side.
(
LMTD = ∆T ) 1
− ( ∆T )
q = (8.01Dc) 2
75.4
ln ) = f (∆T
− 85
)
⎡ ⎤
1
⎢ ⎥ ln 75.4
= (8.01× −9.6 61.3×1.00)(10)(150
⎡ ⎤ −0.1198 = 80.1º F −135) = 73,65
q = (8.01Dc) f (∆T ) 73650
T out
= 73650 ⎢
⎣(∆T ) 2 ⎦
85
⎥ + 50 = 74.6º F
q T= out (8.01Dc) = (8.01× 62.4 ⎣ ×1.00)(6) ⎦
(8.01×
f (∆T
62.4
)
×1.00)(6)
= (8.01× 61.3×1.00)(10)(150 + = 74.6º F −135) = 73,650 Btu
q = (8.01Dc) f (∆T ) = (8.01× 61.3×1.00)(10)(150 −135) = 73,650 Btu
hr
Since both inlet and outlet 73650
hr
temperatures are now known,
T
the log mean out
=
+ 50 = 74.6º F
temperature (8.01× 62.4 ×1.00)(6) difference can be calculated:
1
1
U = 73650
⎡ 1
⎢ + R
h ffh
+ ∆ x
h
k + R + 1 =
= 7
⎤ ⎡
⎤
LMTD = ∆T
out
= 73650( ) 1
− ( ∆T ) 2
75.4 − 85
LMTD = ∆T + 50 = 74.6º F
T ( )
ffc ⎥ ⎢
⎣
h c ⎦
1
0.00167 ft
1
⎥
⎢
+ 0 + + 0 +
⎥
ln (∆T ) =
⎡ ⎤
⎢
1 Btu
Btu
Btu
⎢ 250 ⎥ ln 75.4 = −9.6
1
− ( ∆T ) 2
75.4 − 85
⎡ ⎤ −0.1198 = 80.1º F
ln (∆T ) =
⎡ ⎤
1
⎢ ⎥ ln 75.4 = −9.6
out
(8.01× = 62.4 ×1.00)(6) + 50 = 74.6º F
(8.01× 62.4 ×1.00)(6)
⎡ ⎤ −0.1198 = 80.1º F
( ⎢ ⎢ 29
100
⎥
⎣
⎢ hr • ft 2 •º F hr • ft •º F hr • ft 2 •º F ⎦
⎥
⎣(∆T ) 2 ⎦
⎥
⎣(∆T ) 2 ⎦
85
⎥
⎣ ⎦ ⎣ ⎦
LMTD = ∆T ) − ( ∆T ) 1 2
75.4( − 85
LMTD
ln (∆T ) = = ∆T )
⎡ ⎤
1
⎢
The U value ⎣(∆T ) 2 for ⎦
⎥ ln 75.4 = 1
− ( ∆T ) −9.6
⎡
the ⎣
⎢
⎤ −0.1198 = 80.1º 2
75.4 − 85
F
85
⎥
heat ⎦
( exchanger can also be determined
based LMTD on = ∆T ) 1
− ( ∆Tln ) (∆T ) =
⎡ ⎤
1
⎢
2
75.4⎥
ln 75.4 = −9.6
⎡ ⎤ −0.1198 = 80.1º F
− 85 ⎢
Formula 4-7:
ln (∆T ) ⎣= (∆T )
⎡ ⎤
1
⎢ 1 ⎥
q UA(LMTD) = 73,650 Btu ln 75.4 = −9.6
2 ⎦
85
⎥
⎣ ⎦
⎡ ⎤ −0.1198 = 80.1º F
hr = ⎢
(71.1 Btu
1
U =
1
1
)A(80.1º F)
U =
hr • ft 2 1
1
•º F
⎡ 1
⎢ + R
h ffh
+ ∆ x
h
k R + 1 =
⎤ ⎡
h ffh
+ ∆ x
h
k + R + 1 =
1
Btu
U = ⎣(∆T 1
+ R
⎡
⎤ ⎡
h ffh
+ ∆ x
h ⎢ k + R + 1 =
) 2 ⎦
85
⎥
⎣ ⎦
= 71.1
⎡
⎤ ⎡
⎤ hr • ft 2 •º F
⎢
ffc ⎥ ⎢
⎣
h c ⎦
1
ffc ⎥ 0.00167 ⎢ ft
1
⎥
⎢
+ 0 + +
⎣
h c ⎦
10 +
⎥
0.00167 ft
⎢ Btu
Btu
Btu
250 1 ffc 29 ⎢ ⎥ ⎢ 100 +
⎣
h c ⎦ Btu 1
0 +
⎥
1
0.00167
+ 0 +
U = ⎣
⎢ hr • ft
⎢
ft
2 •º F hr • ft •º F hr • ft
250 ⎢
2 •º F ⎦
⎥
29 +
Btu
1
0 +
+ R hr • ⎢ft Btu
Btu
Notice that the unit for the thickness of the heat exchanger
73,650 Btu
⎣
⎢
250
2 •º F hr • ft •º F
100 29 h
1h ffh
+ ∆ x
h
k + R + 1 =
⎡
⎤ ⎡
⎢
ffc ⎥ ⎢
⎣
h c ⎦
1
⎣
⎢ hr • ft 2 •º 1 0.00167 ft
U =
F hr • ft •º
plates 1was changed from 0.02 inches to 0.00167 ft. This is
A q = UA(LMTD) = 73,650 Btu hr = 12.9 ft 2
hr
necessary to Btu maintain = (71.1 Btu
+ R )A(80.1º F)
h ffh
+ ∆ x
hr • ft
consistent 2 •º F
units in all terms, yielding
(71.1 h
k + R + 1 =
⎢
+ 0 + + 0 +
⎡
⎤ ⎡ ⎢ Btu
Btu
250
29
⎢
ffc ⎥ ⎢
⎣
the units hr of • Btu/hr/ft 2 •º F )(80.1º h c
2 /ºF F) ⎦
1 ⎣
⎢ hr • ft 2 0.00167
•º F
ft
hr • ft •º F
1
⎢
+ 0 + + 0 +
for ⎢the result. Btu
Btu
Btu
250
q = UA(LMTD) = 73,650 Btu
29
100
⎣
⎢ hr
Formula 4-5 can now be set up hr = • (71.1 ft 2 •º F Btu hr • ft •º F hr • ft
and solved for the required
q = UA(LMTD) = 73,650 Btu
)A(80.1º F)
2
73,650 Btu
hr
internal surface area of the heat exchanger.
( A)(LMTD)
hr = • (71.1
ft 2 •º F
A =
hr = 12.9 ft
Btu
2
Btu
hr • ft 2 •º F )A(80.1º F
(71.1
hr • ft 2 •º q F = )(80.1º UA(LMTD) F) = 73,650 Btu
q =
hr = (71.1 Btu
)A(80.1º F)
hr • ft 2 •º F
⎡
⎛
1
+ R ln d ⎞
⎤
o
q ⎢= UA(LMTD) =
ffi ⎝
⎜ 73,650 d
+ i ⎠
⎟
Btu
⎢
i
h i
A i
2π kL + R ffo
+ 1 ⎥
( A)(LMTD)
73,650 Btu
q =
hr = (71.1 Btu
)A(80.1º F)
⎥ hr • ft 2 •º F ⎡
⎛
A o
A o
h ⎥
A =
hr
o
⎥
⎣
Btu
(71.1
Btu = 12.9 ft 2
1
+ R ln d ⎞
⎤
o
⎢
ffi ⎝
⎜ d
+ i ⎠
⎟
⎢
A ⎦⎥
hr • ft 2 •º F )(80.1º hr
F) i
h i
A i
2π kL + R ffo
+ 1 ⎥
⎥
⎢
A o
A o
h ⎥
⎢
73,650 Btu
o
⎥
⎣⎢
⎦⎥
A =
hr = 12.9 ft 2
q = (8.01Dc) f (∆T )
73,650 Btu
= 12.9 ft 2
Btu
(71.1
2 •º F) A =
hr • ft 2 •º F )(80.1º F) = 12.9 ft 2
Btu
(71.1 ( A)(LMTD)
q = hr • ft 2 •º F )(80.1º F)
⎡
⎛
1
q = (8.01Dc) f (∆T ) = (8.01× 61.3×1.00)(10)(150 −135) = 73,650 Btu ( A)(LMTD)
q A=
+ R ln d ⎞
⎤
o
⎢
ffi ⎝
⎜ d
( A)(LMTD)
+ i ⎠
⎟
⎢
39
i
h i ⎡ A i
2π kL hr
+ R ffo
+ 1 ⎥
q =
⎥
⎢ ⎡
⎛
⎛
A
ln d o ⎞
A o
h ⎥
⎢ ( A)(LMTD)
o ⎤
o ⎥
q =
⎣⎢
1
1 R ffi ⎝
⎜ d i ⎠
⎟ ⎥
⎡
⎛
+ R ln d ⎞
⎤
o
⎢
ffi ⎝
⎜ d
R ⎦⎥
A d o
⎞
+ i ⎠
⎟
⎤
i
h i
A i
2π kL + R ffo
+ 1 ⎥
⎢
⎥
⎢
A o ffo A o
h o
1⎥
⎥
q = (8.01Dc) f (∆T )
The designer would now consult manufacturer’s
specifications to locate candidate heat exchangers with
internal surface areas 73650 of approximately 12.9 ft 2 .
T out
=
+ 50 = 74.6º F
(8.01× 62.4 ×1.00)(6)
OVERALL RATE OF HEAT TRANSFER –
PIPE HEAT EXCHANGER
Another common “geometry” for heat exchangers involves
one or more cylindrical ( pipes separating the two fluids
LMTD = ∆T ) 1
− ( ∆T ) 2
75.4 − 85
exchanging heat. This is illustrated in Figure 4-10.
ln (∆T ) =
⎡ ⎤
1
⎢ ⎥ ln 75.4 = −9.6
⎡ ⎤ −0.1198 = 80.1º F
⎢
⎣(∆T ) 2 ⎦ ⎣ 85
⎥
⎦
Figure 4-10
q = (8.01Dc) f (∆T ) = (8.01× 61.3×1.00)(10)(150 −135) = 73,650 Btu
hr
do
di
73,650 Btu
A =
hr = 12.9 ft 2
Formula 4-9, which is Btu
(71.1
similar in mathematical form to
Formula 4-6, can be hr • used ft 2 •º F to )(80.1º calculate F) the rate of heat
transfer for a pipe wall separating a hot and cool fluid.
Where:
q = rate of heat transfer across the heat exchanger (Btu/hr)
A i = internal surface area of the pipe (ft 2 )
A o = outside surface area of the pipe (ft 2 )
LMTD = log mean temperature difference established by
the current conditions (ºF)
h i = convection coefficient on the inside of the heat
exchanger (Btu/hr/ft 2 /ºF)
R ffi = thermal resistance of fouling factor on inside of the
pipe (ºF•hr•ft 2 /Btu)
d 0 = outer diameter of the pipe (ft)
d i = inner diameter of the pipe (ft)
L = length of the pipe (ft)
k = thermal conductivity of metal wall of pipe (Btu/hr/ft/ºF)
R ffo = thermal resistance of fouling factor on cool side
(ºF•hr•ft 2 /Btu)
h o = convection coefficient on the cool side of the heat
exchanger (Btu/hr/ft 2 /ºF)
pi = 3.141592654
ln( ) = natural logarithm function of quantity inside ( )
Much of the effort in using a formula such as this involves
gathering information and making estimates for quantities
such as the two convection coefficients (h 1 and h 2 ) and the
two fouling factors (R ffi and R ffo ).
An important point regarding the overall heat transfer
1
1 coefficient is that it depends on Btu
U =
the convection coefficients on
1
+ R both sides of the heat exchanger. If one of these convection
h ffh
+ ∆ x
h
k + R + 1 =
= 71.1
⎡
⎤ ⎡
⎤ hr • ft 2 •º F
⎢
ffc ⎥ ⎢
⎣
h c ⎦
1
0.00167 ftcoefficients 1
⎥
⎢
+ 0 + + 0 + is small ⎥ relative to the other, it will be the factor
⎢ Btu
Btu
Btu
250
29 that limits 100 the overall ⎥ performance of the heat exchanger.
⎣
⎢ hr • ft 2 •º F hr • ft •º F hr • ft 2 •º F ⎦
⎥
outer
fluid
inner
fluid
heat
transfer
q = UA(LMTD) = 73,650 Btu
hr = (71.1
Btu
inner boundary layer
inner fouling hr • ftfilm
2 •º F
pipe wall
outer fouling film
outer boundary layer
Formula 4-9:
( A)(LMTD)
q =
⎡
⎛
1
+ R ln d ⎞
⎤
o
⎢
ffi ⎝
⎜ d
+ i ⎠
⎟
A i
h i
A i
2π kL + R ffo
+ 1 ⎥
⎢
⎥
⎢
A o
A o
h ⎥
⎢
o
⎥
⎣⎢
⎦⎥
)A(80.1º F)
A good example of this is a coil heat exchanger inside a
thermal storage tank. Based on some assumptions for
typical operating conditions, the convection coefficient
governing natural convection on the outer surface of the coil
is only about 6% of the convection coefficient associated
with forced flow inside the coil. This relatively low value
will severely limit the overall heat transfer coefficient and
thus require a substantially larger surface area (A) to create
a given rate of heat transfer, relative to a situation where
forced convection was present in both flow streams.
The difference between the relatively low convection
coefficients associated with natural convection versus the
higher values associated with forced convection can also
be seen in a comparison between the length of finnedtube
baseboard required for a certain heat output, versus
the surface area of a coil in a wall-mounted fan-coil for
equivalent heat output.
One commercially available fan-coil measuring 32 inches
wide and 24 inches tall, when operating at 1.5 gpm, highspeed
fan, and 110ºF average water temperature, can
release approximately 6,700 Btu/hr. Standard residential
finned-tube baseboard, operating at the same average
water temperature and flow rate, can releases about 118
Btu/hr per foot of finned-tube length. To match the output
of the fan-coil, the baseboard’s finned-tube length would
need to be 6700/118 = 57 feet. Figure 4-11 shows a
scaled comparison of these two heat emitters.
Even though the forced flow occurring on the inside of the
copper tube in the baseboard may have a high convection
coefficient, heat transfer is severely limited by the natural
convection occurring between the outer surface of the
finned-tube element and the room air.
40
Figure 4-11
FINNED-TUBE BASEBOARD
output = 6700 Btu/hr
average water temperature = 110 ºF
room air temperature = 68ºF
natural convection on air side
24"
32"
FAN-COIL
output = 6700 Btu/hr
average water temperature = 110 ºF
room air temperature = 68ºF
forced convection on air side
57'-0"
Courtesy of Myson
Courtesy of Nick Feriola @nickdaplumber Instagram
DETERMINING OVERALL RATE OF HEAT TRANSFER
(EFFECTIVENESS METHOD)
The method of analyzing heat exchanger performance
based on LMTD is useful when the entering and leaving
temperatures of both flow streams are known, or they can
be calculated based on an energy balance across the heat
exchanger. However, when there is insufficient information
to calculate the LMTD, the solution must resort to iterative
calculations. These involve making “guesstimates” for the
unknown temperatures, calculating heat transfer rates based
on these temperatures, finding the “error” between the
calculated rates of heat transfer on the two sides of the heat
exchanger, correcting the guesstimated temperatures, and
repeating the procedure. This is possible but very tedious. In
these situations, another method of heat exchanger analysis
is usually preferred. That method is based on the concept of
“effectiveness” of the heat exchanger.
The effectiveness of a heat exchanger is defined by Formula
4-10.
Formula 4-10:
ε = q actual
q max
Where:
e = effectiveness of the heat exchanger (decimal number
between 0 and 1)
q actual = Actual rate of heat transfer across the heat
exchanger (Btu/hr)
q max = Maximum possible rate of heat transfer across the
heat exchanger given the operating conditions (Btu/hr)
Effectiveness is based on the concept that, with a
hypothetical heat exchanger having an infinite internal
area, one of the two fluid streams could undergo the
maximum possible temperature change, from the inlet
temperature of the cold fluid up to the inlet temperature of
the hot fluid. Which fluid could undergo this temperature
change depends on which has the smaller capacitance
rate. The latter term, capacitance rate, is the product of
specific heat times mass flow rate, expressed in units of
Btu/hr/ºF.
For example: Consider a situation where one side of a heat
exchanger operates with water and a flow rate of 5 gpm.
The other side operates with a 50% solution of propylene
glycol and a flow rate of 6 gpm. The average temperature
of the fluids is 120ºF.
⎛
C water
= 5 gallon ⎞ ⎛
⎝
⎜
minute⎠
⎟ 1 Btu ⎞ ⎛ 8.33lb ⎞ ⎛
⎝
⎜
lb•º F ⎠
⎟
⎝
⎜
gallon⎠
⎟
⎝
⎜
60minute
hour
⎞
⎠
⎟
= 2499
Btu
hr⋅º F
41
C 50% PG
= 6 gallon Btu
⎝
⎜
minute⎠
⎟ 0.88
⎝
⎜
lb•º F ⎠
⎟
⎝
⎜
ε = q actual
⎛
q max
C water
= 5 gallon ⎞ ⎛
The capacitance rate for the water side is:
One is called the capacitance ⎝
⎜
minuterate ⎠
⎟ 1 Btu ⎞ ⎛ 8.33lb ⎞
⎝
⎜
ratio. lb•º FIt’s ⎠
⎟
⎝
⎜
simply gallonthe
⎠
⎟
ε = q actual
minimum fluid capacitance rate of the two fluid streams
q
⎛ max
divided by the
q
larger
= ε ( Cfluid capacitance rate, stated as
C water
= 5 gallon ⎞ ⎛
⎝
⎜
minute⎠
⎟ 1 Btu ⎞ ⎛ 8.33lb ⎞ ⎛ 60minute⎞
Btu
min )( T hotin
− T coldin )
⎝
⎜
lb•º F ⎠
⎟
⎝
⎜
gallon⎠
⎟
⎝
⎜
hour ⎠
⎟ = 2499
hr⋅º F Formula 4-12:
⎛
C Formula 4-12:
water
= 5 gallon ⎞ ⎛
The capacitance ⎝
⎜
minute⎠
⎟ 1 Btu ⎞ ⎛ 8.33lb ⎞ ⎛ 60minute⎞
Btu
⎛
⎝
⎜
rate lb•º Ffor ⎠
⎟
⎝
⎜
the gallon side ⎠
⎟
⎝
⎜
operating hour ⎠
⎟ = 2499 C
with propylene hr⋅º F
50% PG
= 6 gallon ⎞ ⎛ Btu ⎞ ⎛
⎝
⎜
glycol is: ⎛
C 50% PG
= 6 gallon ⎞ ⎛ Btu ⎞ ⎛ 8.54lb ⎞ ⎛ 60minute⎞
Btu
⎝
⎜
minute⎠
⎟ 0.88
⎝
⎜
lb•º F ⎠
⎟
⎝
⎜
gallon⎠
⎟
⎝
⎜
hour ⎠
⎟ = 2705 C
hr⋅º F
ratio
= C minute⎠
⎟ 0.88
⎝
⎜
lb•º F ⎠
⎟
⎝
⎜
min
C max
⎛
C 50% PG
= 6 gallon ⎞ ⎛ Btu ⎞ ⎛ 8.54lb ⎞ ⎛ 60minute⎞
Btu
⎝
⎜
minute⎠
⎟ 0.88
⎝
⎜
lb•º F ⎠
⎟
⎝
⎜
gallon⎠
⎟
⎝
⎜
hour ⎠
⎟ = 2705
hr⋅º F Where:
C ratio = capacitance rate ratio (unitless)
q = ε
C min = smaller of the two capacitance rates for the two
( C
Of these two
min )( T
capacitance
hotin
− T coldin )
q = ε ( C
rates, the one associated with flow streams involved NTU (Btu/hr/ºF) = UA min )( T hotin
− T coldin )
the q water = ε ( C min
side )( T hotin
is −smaller T coldin ) and will be designated as C min . C max = larger of the two capacitance C min rates for the two flow
The larger capacitance rate will be designated C max . streams involved (Btu/hr/ºF)
C
It’s important ratio
= C min
max
to note that the physical properties of the The other quantity that’s needed is called NTU, which
two Cfluids ratio
= C min
were ε = referenced q actual
to an estimated average stands for number C of transfer 1− e ( − NTU
units, )(1−C ratio
and )
ratio
= C min
is defined as
C
temperature max
in the qheat max exchanger. Also notice that the Formula 4-13. ε = C max
units for capacitance rate are Btu/hr/ºF. One can interpret
1− (C
the value NTU = of UA
ratio
)e ( − NTU )(1−C ratio )
Ccapacitance rate as the rate of heat transfer Formula 4-13:
min
(in Btu/hr) NTU = that UA each fluid stream could carry into or out of
C
the heat exchanger min
per
⎛
degree F change in temperature
of that stream. C
1− e water ( = 5 gallon ⎞ ⎛
− NTU )(1−C ratio )
ε =
The fluid 1− 1−
having (Ce ( ⎝
⎜
minute⎠
⎟ 1 Btu ⎞ ⎛ 8.33lb ⎞ ⎛ 60minute⎞
NTU Btu = UA
⎝
⎜
lb•º F ⎠
⎟
⎝
⎜
gallon⎠
⎟
⎝
⎜
hour ⎠
⎟ = 2499
hr⋅º F
− NTU
ratio
)ethe ()(1−C − NTU ratio
lower )(1−C )
ratio ε =
NTU C min
ε =
)
of the two capacitance rates is Where:
1− (C
one of the factors ratio
)e ( 1+ NTU
− NTU
that )(1−C ratio )
limit the maximum possible rate of NTU = number of transfer units (unitless number)
heat transfer. That maximum possible rate for a hypothetical U = overall heat transfer coefficient for the heat exchanger
1− e ( − NTU )(1−C ratio )
infinitely large heat exchanger would be the lower of the (Btu/hr/ft 2 /ºF)
two capacitance NTU rates (e.g., C min ), multiplied by the inlet A = internal area of the heat exchanger (ft 2 )
ε = 1+ NTU NTU C
temperature 1+ NTUof 50%
the PG
= 6 gallon
ε =
⎛ ⎞ ⎛ Btu ⎞ ⎛ 8.54lb ⎞ ⎛ 60minute⎞
hot ⎝
⎜fluid minute minus ⎠
⎟ 0.88
the ⎝
⎜ inlet lb•º temperature F ⎠
⎟
⎝
⎜
gallon of ⎠
⎟
⎝
⎜C min hour = smaller ⎠
⎟ = 27051− Btu (C ratio
)e ( − NTU )(1−C ratio )
of the hr⋅º two capacitance rates of the two
⎛ F
the cold fluid. This allows the rate of heat transfer across flow streams (Btu/hr/ºF) C water
= 5 gallon ⎞ ⎛
the heat exchanger to be written as Formula 4-11.
⎝
⎜
minute⎠
⎟ 1 Btu ⎞ ⎛
⎝
⎜
lb•º F ⎠
⎟
⎝
⎜
One can think of the NTU as the ratio of the physical ability
⎛
Formula 4-11:
⎛
⎞ ⎛
of a heat exchanger to transfer heat (based on its internal
⎝
⎜
⎠
⎟ 1 Btu ⎞ ⎛ 8.33lb ⎞ ⎛ 60minute⎞
Btu
C water
= 5 gallon ⎞ ⎛
⎝
⎜
lb•º F ⎠
⎟
⎝
⎜
gallon⎠
⎟
⎝
⎜
hour ⎠
⎟ = 2499
⎝
⎜
minute⎠
⎟ 1 Btu ⎞ ⎛ 8.33lb ⎞ ⎛ 60minute⎞
Btu
⎝
⎜
lb•º F ⎠
⎟
⎝
⎜
gallon⎠
⎟
⎝
⎜
hour ⎠
⎟ = 2499
hr⋅º Fhr⋅º F
area and the overall ε =
NTU
heat transfer coefficient) divided by
1+ NTU
q = ε ( C min )( T hotin
− T coldin )
the ability of the “least able” flow stream to convey heat
into or out of the heat exchanger. For example, a heat
Where:
exchanger with a large internal area, and operating under
q = actual rate of heat exchange across the heat exchanger favorable convection conditions, but also operating with a
(Btu/hr)
low capacitance rate on one side, would have high NTU.
e = effectiveness of the heat exchanger (unitless value Because of its large internal area and high convection
between 0 and 1) C ratio
= C ⎛
min
C
coefficients. it would water
= 5 gallon ⎞ ⎛
be very “effective” in transferring
C ⎝
⎜
minute⎠
⎟ 1 Btu ⎞ ⎛
⎝
⎜
lb•º F ⎠
⎟
⎝
⎜
C min = the smaller of the two max fluid capacitance rates (Btu/ heat to (or from) the fluid with the lower capacitance
hr/ºF)
rate, possibly even approaching the performance of a
T hotin = inlet temperature of the hot fluid (ºF)
hypothetical infinite heat exchanger, which would have an
T coldin = inlet temperature of the cold fluid (ºF)
effectiveness of 1.0.
NTU = UA
Although Formula 4-11 appears C relatively simple, there Once the capacitance rate ratio (C ratio ) and number of
min
is additional work involved to determine the value of transfer units (NTU) are calculated, the effectiveness
effectiveness (e).
for a specific type of heat exchanger can be calculated
Two additional quantities need to be defined in order to based on either formulas or graphs in heat transfer
determine the value of effectiveness (e).
reference books.
1− e ( − NTU )(1−C ratio )
ε =
( )(1−C ratio )
42
1− (C ratio
)e − NTU
q max
⎛
⎞ ⎛
⎞ ⎛ 8.54lb ⎞ ⎛
gallon⎠
⎟
⎝
⎜
⎛
⎝
⎜
8.33lb ⎞ ⎛
gallon⎠
⎟
⎝
⎜
60mi
ho
60minute
hour
8.54lb ⎞ ⎛
gallon⎠
⎟
⎝
⎜
8.33lb ⎞ ⎛
gallon⎠
⎟
⎝
⎜
60m
ho
60min
hou
60min
hou
C
C ratio
= C ratio min= C min
C max
C max
For example, for any counterflow heat exchanger, the
effectiveness can be calculated using Formula 4-14a or
4-14b. Formula 4-14a is used in the typical case where
ε C= q ratio < 1. Formula NTU 4-14b is used in the less typical but
possible case
NTU
where
= UA = UA
actual
q max
ratio
C= 1 (e.g., both fluid streams
C min
have the same capacitance min
rate ratio.
Formula ⎛
C
4-14a: (use when Cratio ( <1)
water
= 5 gallon ⎞ ⎛
⎝
⎜
minute⎠
⎟ 1 Btu ⎞ ⎛ 8.33lb ⎞ ⎛ 60minute⎞
Btu
⎝
⎜
lb•º F ⎠
⎟
⎝
⎜
gallon⎠
⎟
⎝
⎜
hour
)(1−C ⎠
⎟ =
ratio ) 2499
hr⋅º F
1− e ( − 1− NTU e − NTU
)(1−C ratio
ε =
)
ε = 1− (C
1− (C ratio
)e − ratio NTU )e − NTU
⎛
C 50% PG
= 6 gallon ⎞ ⎛ Btu ⎞ ⎛ 8.54lb ⎞ ⎛
Formula ⎝
⎜
minute 4-14b: ⎠
⎟ 0.88
⎝
⎜
(use when lb•º FCratio ⎠
⎟
⎝
⎜
gallon = 1) ⎠
⎟
⎝
⎜
ε =
NTU ε =
NTU
1+ NTU
1+ NTU
( )( T hotin
− T coldin )
( ( )(1−C ratio )
)(1−C ratio )
60minute
hour
⎞
⎠
⎟
= 2705
Btu
hr⋅º F
Figure 4-12
effectiveness
vs. NTU
for counterflow
heat exchanger
q = ε C min
0.4
The terms shown in red in Formula 4-14a are identical and
thus only have to be calculated once.
0.3
⎛
0.2
The effectiveness of a counterflow heat exchanger as
C
a function of the number of transfer units (NTU) and the
0.1
capacity rate C water ratio = (C min
5 gallon ⎛
C ⎛ ⎞
/C max ) can also be found using
Figure 4-12.
⎝
⎜
minute⎠
⎟ 1 Btu
water
= 5 gallon ⎞
⎛ ⎞ ⎛ 8.33lb ⎞ ⎛ 60minute⎞
Btu
⎝
⎜
minute
⎝
⎜ ⎠
⎟ 1 Btu
ratio
= C C 50% PG
= 6 gallon ⎞ ⎛ Btu ⎞ ⎛ 8.54lb ⎞ ⎛ 60minute⎞
⎝
⎜
minute⎠
⎟ 0.88
⎝
⎜
lb•º F ⎠
⎟
⎝
⎜
gallon⎠
⎟
⎝
⎜
hour ⎠
⎟ =
min
⎛ ⎞ ⎛ 8.33lb ⎞ ⎛ 60minute⎞
Btu
C
⎛
max
⎝
⎜
lb•º
lb•º F ⎠
⎟
⎝
⎜ F ⎠
⎟
⎝
⎜
gallon
gallon⎠
⎟
⎝
⎜ ⎠
⎟
⎝
⎜C 50%
hour
PG
= 6
hour ⎠
⎟ = 2499 gallon ⎞ ⎛ Btu ⎞ ⎛ 8.54lb ⎞ ⎛ 60minute⎞
Btu
⎠
⎟ = 2499
⎝
⎜
minute⎠
⎟ 0.88
⎝
⎜
hr⋅º lb•º FF⎠
⎟
⎝
⎜
gallon⎠
⎟
⎝
⎜
hour ⎠
⎟ = 2705
⎛
hr⋅º F
C 50% PG
= 60
gallon ⎞ ⎛ Btu ⎞ ⎛ 8.54lb ⎞ ⎛ 60minute⎞
Btu
⎝
⎜
minutehr⋅º ⎠
⎟ 0.88
⎝
⎜ F
lb•º F ⎠
⎟
⎝
⎜
gallon⎠
⎟
⎝
⎜
hour ⎠
⎟ = 2705
hr⋅º F
0 1 2 3 4 5
Here’s NTU = an UA
example: Consider a counterflow heat exchanger
NTU
C
with an internal
min
area of 20 ft 2 C
. It is operating with the two
ratio
= C min
= 2499
C max
2705 = 0.924
previously described flow streams of water at 5 gpm and C ratio
= C min
= 2499
50% propylene glycol at 6 gpm. The water stream enters The number max
2705 = 0.924
of transfer units is:
at ε = 150ºF,
1− e
and ( C − NTU
the )(1−C ratio
propylene )
ratio
= C min
= 2499
C
glycol solution enters at 60ºF.
max
2705 = 0.924
An 1− analysis (C ratio
)e ( of − NTU
the )(1−C ratio
convection )
coefficients and fouling
NTU = UA (
= 150 )20
factors on both sides of the heat exchanger has been
C min
2499 = 1.2
NTU = UA (
= 150 )20
used to establish the overall heat transfer coefficient (U)
C min
2499 = 1.2
as 150 Btu/hr/ft 2 NTU = UA (
= 150 )20
/ºF. Determine the rate of heat transfer These values can now be entered into Formula 4-14a:
across ε =
NTU
C min
2499 = 1.2
the heat exchanger.
1− e ( − NTU )(1−C ratio )
1+ NTU
ε =
(C
The two fluid capacitance rates have already been
ratio
)e ( = 1− e ( −1.2)(1−0.924)
1− e ( − NTU )(1−C ratio ) 1− (0.924)e ( = 0.0872
− NTU )(1−C ratio )
ε =
−1.2)(1−0.924)
0.1565
1− (C ratio
)e ( = 1− e ( −1.2)(1−0.924)
0.0872
− NTU )(1−C ratio ) (
=
1− (0.924)e )(1−0.924)
1− e ( 0.1565 = 0.557
− NTU )(1−C ratio )
ε =
determined as:
1− (C ratio
)e ( = 1− e ( −1.2)(1−0.924)
0.0872
− NTU )(1−C ratio ) (
=
1− (0.924)e −1.2 )(1−0.924)
0.1565 = 0.557
⎛
The final step is to apply Formula 4-11:
C water
= 5 gallon ⎞ ⎛
⎝
⎜
minute⎠
⎟ 1 Btu ⎞ ⎛ 8.33lb ⎞ ⎛ 60minute⎞
Btu
⎝
⎜
lb•º F ⎠
⎟
⎝
⎜
gallon⎠
⎟
⎝
⎜
hour ⎠
⎟ = 2499
hr⋅º F
q = ε ( C min )( T hotin
q −= T coldin
ε ( C) min
= ) 0.557 ( T hotin
( 2499 − T coldin
)( 150 ) = −0.557 60) = ( 125,280 2499) ( 150 Btu − 60) = 125,28
hr
q = ε ( C
⎛
C 50% PG
= 6 gallon
min )( T hotin
− T coldin ) = 0.557( 2499) ( 150 − 60) = 125,280 Btu
⎞ ⎛ Btu ⎞ ⎛ 8.54lb ⎞ ⎛ 60minute⎞
Btu
hr
⎝
⎜
minute⎠
⎟ 0.88
⎝
⎜
lb•º F ⎠
⎟
⎝
⎜
gallon⎠
⎟
⎝
⎜
hour ⎠
⎟ = 2705
⎛
hr⋅º F
C 50% PG
= 6 gallon ⎞ ⎛ Btu ⎞ ⎛ 8.54lb ⎞ ⎛ 60minute⎞
Btu
⎝
⎜
minute⎠
⎟ 0.88
⎝
⎜
lb•º F ⎠
⎟
⎝
⎜
gallon⎠
⎟
⎝
⎜
hour ⎠
⎟ = 2705
hr⋅º F
θ =
Of these two capacitance rates, the one for the water side SOFTWARE-BASED ∆T
LMTD θ =
∆T HEAT EXCHANGER SELECTION
is smaller. Thus, the capacitance rate ratio is:
The procedures introduced LMTD thus far — the LMTD method,
C ratio
= C min
= 2499
θ =
∆T
and the effective (e) method —can be used for manual
C max
2705 = 0.924
LMTD
calculations of heat exchanger performance. Both of
C ratio
= C min
= 2499
these approaches demonstrate that the mathematics
C max
2705 = 0.924
⎡ ∆T
q 2
= q 2
1
necessary ∆T for sizing heat ⎡ ∆T exchangers can be complex. One
NTU = UA (
= 150
⎣
⎢
⎤
⎡ ∆T 2 ⎦
⎥
q
)20
of the more complex 2
= q 2
⎤
1 ⎢
tasks
⎥
—that was not documented in
C min
2499 = 1.2
q 2
= q 2
⎤
1 ⎢ ⎥
⎣ ∆T
⎣ ∆T 2 ⎦
2 ⎦
NTU = UA (
= 150 )20
1− e ( − NTU )(1−C ratio ) C 43
ε =
1− (C ratio
)e ( = 1− e ( −1.2)(1−0.924)
min
2499 = 1.2 ⎡ ∆T
q 2
= q 2
⎤ ⎡140 − 65⎤
Btu
1 ⎢ ⎥ = 5,000
0.0872
∆T
⎢
− NTU )(1−C ratio ) (
=
1− (0.924)e −1.2 )(1−0.924)
0.1565 = 0.557
⎣ 1 ⎦ ⎣180 65
⎥ = 3,261
⎡ ∆T ∆T
q 2
= q 2
⎤ ⎦ hr
q
⎡140 − 65⎤
Btu
2
= q 2
⎤ ⎡140 − 65⎤
Btu
1 ⎢ ⎥ = 5,000
1 ⎥ = 5,000 ∆T
⎢
1 ⎦ ⎣180 − 65
⎥ = 3,261
⎣ ∆T
⎢
1 ⎦ ⎣180 − 65
⎥ = 3,261
⎦ hr
⎦ hr
effectiveness (e)
1
0.9
0.8
0.7
0.6
0.5
Cmin/Cmax=0.2
Cmin/Cmax=0.4
Cmin/Cmax=0.6
Cmin/Cmax=0.8
Cmin/Cmax=1.0
Figure 4-13a
Source: Kelvion, www.flatplateselect.com
Figure 4-13b
44
Figure 4-13c
the minimum surface area required for the specified heat
exchange rate. In some cases, the extra surface area is
necessary to limit the pressure drop of one or both sides of
the heat exchanger. The software provides performance
data such as the overall heat transfer coefficient, flow
velocity and rate of heat transfer. Physical data on
the candidate heat exchangers, including volume and
dimensions, are also listed. The ability of this software
to evaluate several “what if” scenarios in a fraction of
the time required for manual calculations makes it an
essential design tool.
HEAT EXCHANGER PERFORMANCE INDICES
Two indices are commonly used to describe the actual
performance, or performance target, of heat exchangers.
They are:
• Approach temperature difference
• Thermal length
Approach temperature difference is the temperature of the
fluid entering the hot side of the heat exchanger, minus the
temperature of the “cooler” fluid leaving the heat exchanger.
Figure 4-14 illustrates this difference.
the previous examples — is estimating the values for the
convection coefficients (h h and h c ) in Formulas 4-7 and 4-9.
Those calculations are often iterative and time consuming.
As such, they don’t lend themselves to repeated “what if”
scenarios when a designer is trying to select an appropriate
heat exchanger.
Many manufacturers now offer software that can rapidly
evaluate the necessary mathematics for sizing and
selection of their heat exchangers. In some cases, this
software is accessed online. In other cases, it’s necessary
to download and install the software. Some suppliers also
require submittal of design parameters and operate the
software in-house.
A theoretically infinite heat exchanger would have an
approach temperature difference of 0. Any real heat
exchanger will have an approach temperature difference
greater than zero. Some brazed plate and plate & frame
heat exchangers can operate at approach temperature
differences as low as approximately 2ºF, although this
Figure 4-14
Approach
temperature =
difference
T hot in
( )
T hot in
- T cool out
T cool out
One example of online-based sizing and selection software
is FlatPlateSELECT, available at http://flatplateselect.
com/site/hx/chooseapp.aspx. Figures 4-13a, b and c
show some examples of this software.
Some heat exchanger sizing and selection software
produces a list of several specific models that can meet
the requirements based on input data. They also show an
“oversurface” percent, which is the percent by which the
surface area in the suggested heat exchanger exceeds
T hot out
T cool in
45
NTU = UA (
= 150 )20
C min
2499 = 1.2
requires a relatively large heat transfer area, and thus
a large heat exchanger. The expense of such a large
heat exchanger may not be justified based solely on
a very small approach temperature difference. More
typical approach temperature differences, with the heat
exchanger operating at design load conditions, range
from 4 to 10ºF.
Approach temperature differences are often selected
based on the characteristics of the heat source. For
example, the coefficient of performance (COP) of waterto-water
and air-to-water heat pumps is very dependent
on the load water temperature (e.g., the water temperature
leaving the heat pump’s compressor). If the heat produced
by a heat pump needs to pass through a heat exchanger,
the approach temperature difference should be as low
as practically and economically justifiable. A suggested
maximum approach temperature difference for heat pump
applications is 5ºF.
In another scenario, where a boiler operating at 180ºF is
transferring heat to a load operating at 110ºF, the approach
temperature difference is able to be much higher without
significantly affecting the boiler’s efficiency. This allows
use of a much smaller and less expensive heat exchanger.
1− e ( − NTU )(1−C ratio )
ε =
Keep in mind that approach 1− (C temperature difference is based
on a specific steady-state load. ratio
)e ( = 1− e −1.2
− NTU )(1−C ratio ) 1− (0.924)e −1.2
The measured approach
temperature difference of a heat exchanger can vary widely
when the system is operating in transient conditions.
THERMAL LENGTH OF
q =
A εHEAT C
EXCHANGER
Another concept that helps min
describe the expected duty of
a heat exchanger is called thermal length. It’s defined by
Formula 4-15.
Formula 4-15:
θ =
Where:
q = (Greek letter theta) = thermal length (unitless)
∆T = temperature change ⎡of ∆Tfluid (inlet to outlet) on one
q
side of the heat exchanger 2
= q 2
⎤
1 ⎢(ºF)
⎥
⎣ ∆T 2 ⎦
LMTD = log mean temperature drop at which the heat
exchanger is operating (ºF)
( )(1−0.924)
( )(1−0.92
( )( T hotin
− T coldin ) = 0.557( 2499) ( 150 − 6
∆T
LMTD
Low values of (q) indicate heat transfer conditions that are
⎡ ∆T
relatively easy to accomplish. q High values of (q) represent
2
= q 2
⎤ ⎡140 − 65⎤
Btu
1 ⎢ ⎥ = 5,000
situations where there is very ∆T
⎢
⎣ 1 ⎦little difference 180 − 65
⎥ = 3,261
⎣ between ⎦ hr
the temperature of the fluid providing heat and the fluid
Figure 4-15
T hotin
low thermal length
θ =
∆T
LMTD = 10
12 = 0.83
T hotin
high thermal length
θ =
∆T
LMTD = 10 2 = 5.0
T cold
out
T cold
out
hot fluid
cold fluid
LMTD = 12ºF
T hot
out
∆T = 10ºF
hot fluid
cold fluid
LMTD = 2ºF
T hot
out
∆T = 10ºF
Tcold
in
T hotin
Tcold
in
T hotin
T hot
out
T hot
out
T cold
out
counterflow
Tcold
in
T cold
out
counterflow
Tcold
in
46
absorbing heat. The latter is a much more challenging
condition and generally requires flat plate heat exchangers
with longer plates. Figure 4-15 illustrates the difference
between heat transfer requirements having low and high
thermal lengths (q).
Heat transfer theory limits the thermal length (q) of a single
shell & tube heat exchanger to 1.0. Higher values of (q) are
possible if multiple shell & tube heat exchangers are piped
in series, which is uncommon. Flat plate heat exchangers
can operate with thermal lengths up to approximately
10.0. This makes flat plate heat exchangers better suited
to situations where the temperature differences between
the fluids is minimal. The thermal length of flat plate heat
exchangers can also be affected by different pressing
patterns on the plate. Manufacturers can advise on
optimal plate patterns based on the intended duty of the
heat exchanger.
THERMAL PERFORMANCE OF
WATER-TO-AIR HEAT EXCHANGERS
There are several types of heat emitters that transfer
heat from a stream of water to a fan-forced stream of air.
Section 2 described several of these devices, categorizing
them as fan-coils or air handlers. This section will describe
thermal performance concepts for these devices.
The heat output of a fan-coil or air handler is dependent
on several factors. Some are fixed by the design of the
product and cannot be changed. These include the:
• Surface area of the coil
• Size, spacing and thickness of the fins
• Number of tube passes through the fins
• Air-moving ability of the blower
Other performance factors depend on the system into
which the fan-coil is installed and the conditions under
which that system operates. These include the:
• Entering water temperature
• Water flow rate through the coil
• Entering air temperature
• Air flow rate through the coil
Most manufactures publish tables or graphs showing
the heat output of a fan-coil or air handler over a range
of selected operating conditions, such as entering water
temperature, entering room air temperature and the
flow rates of both the water and air streams. However,
because heat output varies continuously with changes
in temperatures or flow rates, there is no guarantee that
the unit will operate at one of the conditions listed by the
manufacturer.
Heat output (Btu/hr)
Figure 4-16
6000
5000
4000
3000
2000
1000
Figure 4-16 shows representative heat output from a small
fan-coil unit as a function 1− ε (C =
ratio of entering )e − NTU
water temperature
1− (C
and different water flow rates. ratio
)e − NTU
The entering room air
temperature is assumed constant at 65ºF, and the fan is
set to its highest speed.
The manner in which the heat output of a fan-coil or
q = ε C
air handler varies with q = min
εoperating ( C conditions or coil
min ) T hotin
− T coldin
configuration can be summarized in four principles:
Principle #1: The heat output of a fan-coil or
air handler is approximately proportional to the
temperature difference between the entering air and
entering water. ∆T
This principle can be expressed mathematically as:
Formula 4-16:
⎛
C50%
PG
= 6 gallon ⎞ ⎛ Btu ⎞ ⎛ 8.54lb ⎞ ⎛ 6
⎝
⎜ ⎛
minute⎠
⎟ 0.88
C ⎝
⎜
lb•º F ⎠
⎟
⎝
⎜
gallon⎠
⎟
⎝
⎜
50% PG
= 6 gallon ⎞ ⎛ Btu ⎞ ⎛ 8.54lb
⎝
⎜
minute⎠
⎟ 0.88
⎝
⎜
lb•º F ⎠
⎟
⎝
⎜
gallon
High speed fan
Entering air temperature = 65ºF
0
100 110 120 130 140
Entering water temperature (ºF)
5.0 gpm
2.5 gpm
1.5 gpm
0.5 gpm
C ratio
= C min
= 2499
CC max
2705 = 0.924
ratio
= C min
= 2499
C max
2705 = 0.924
NTU = UA (
= 150 )20
NTU C min
= UA (
2499 = 150=
)20 1.2
C min
2499 = 1.2
ε =
θ =
( )(1−C ratio )
1− e − NTU
( = 1− e −1.2
1− e ( − NTU )(1−C ratio )
)(1−C
(
ratio ) 1− (0.924)e = 1− ( e ( −1.2)(1
−1.2)(1−0.
)(1−C ratio ) 1− (0.924)e ( −
Where:
q 2 = heat output at an assumed water-to-air temperature
difference (Btu/hr) ⎡ ∆T
q 1 = heat output q 2
= at qa known 2
⎤ ⎡140 − 65⎤
1 ⎢ ⎡⎥ ∆T = water-to-air 5,000 temperature
difference (Btu/hr) q ∆T
⎢
⎣
⎦ 180 − 65
⎥
2
= q 2
⎤ ⎡140 − 65⎤
1 ⎢ ⎥ = 5,000 ⎣ ⎦
∆T
⎢
∆T 1 = difference between entering ⎣ 1 water ⎦ and entering ⎣180 − 65
air ⎦
temperatures at a known operating condition (ºF)
∆T 2 = difference between entering water and entering air
temperatures at the assumed operating condition (ºF)
( )(1−0.924)
( )( T hotin
− T coldin ) = 0.557( 2499) ( 150 −
( ) = 0.557( 2499) ( 1
LMTD θ =
∆T
LMTD
⎡ ∆T
q 2
= q 2
⎤
1 ⎢ ⎡⎥
∆T
q 2 ⎣= ∆Tq 2
⎤
12
⎢⎦
⎥
⎣ ∆T 2 ⎦
Btu
= 3,261
⎥ = 3,261 hr
47
C ratio
= C min
C max
= 2499
2705 = 0.924
Figure 4-17
manufacturer's data
NTU 6000= UA (
= 150 )20
C min
2499 = 1.2
5000
⎡ ∆T
q 2
= q 2
⎤ ⎡140 − 50 ⎤
1 ⎢ ⎥ = 3,261
⎣ ∆T
⎢
1 ⎦ ⎣140 − 65
⎥
⎦
= 3,913
Btu
hr
Although principle #1 is helpful for quick performance
estimates, it should not replace the use of thermal rating
data supplied by the manufacturer when it is available.
Heat output (Btu/hr)
1− e ( − NTU )(1−C ratio )
Principle #2: Increasing the fluid flow rate through
4000
the coil will marginally increase the heating capacity
1− (C ratio
)e ( = 1− e ( −1.2)(1−0.924)
0.0872
− NTU )(1−C ratio ) (
=
1− (0.924)e −1.2 )(1−0.924)
0.1565 = 0.557
of a fan-coil or air handler.
3000
Some heating professionals “instinctively” disagree with
this statement. They reason that because the water moves
2000
through the coil at a faster speed, it has less time during
ε =
q = ε C min
θ =
1000
0
0 20 40 60 80 100 120 140 160
∆T Entering water temp. - air temp. (ºF)
LMTD
( )( T hotin
− T coldin ) = 0.557( 2499) ( 150 − 60) = 125,280 Btu
hr
Figure 4-18
constant air
flow rate
For example, assume a fan-coil has a known output of 5,000
Btu/hr when supplied with 180ºF water and operating with
room air entering ⎡ ∆T at 65ºF. The designer wants to estimate
the qfan-coil’s 2
= q 2
⎤
1 ⎢ output ⎥ when supplied with 140ºF water, with
all other conditions ⎣ ∆T 2 ⎦ (e.g., room air temperature, water flow
rate and air flow rate) remaining the same. Formula 4-16
yields the estimated output:
constant
inlet air
temperature
constant
inlet water
temperature
⎡ ∆T
q 2
= q 2
⎤ ⎡140 − 65⎤
1 ⎢ ⎥ = 5,000
⎣ ∆T
⎢
1 ⎦ ⎣180 − 65
⎥
⎦
= 3,261
Btu
hr
A proportional relationship between any two quantities will
result in a straight line that passes through or very close
to the origin when the data is plotted on a graph. Figure
4-17 shows this to be true when the heat output data for a
small fan-coil is plotted against the difference between the
entering water and entering air temperatures.
This relationship can also be used to estimate the heat
output from a fan-coil or air handler when the entering
air temperature is above or below normal comfort
temperatures. For example, if a fan-coil can deliver 3,261
Btu/hr with an entering water temperature of 140°F and
entering air temperature of 65°F, its output while heating
a garage maintained at 50°F could be estimated using the
same formula:
Heat output (Btu/hr)
4500
4000
10
9
3500
8
3000
heat output 7
2500
6
2000
∆T
5
4
1500
3
1000
2
500
1
0
0
0 0.5 1 1.5 2 2.5 3 3.5 4 4.5 5
water flow rate (gpm)
∆T (ºF)
48
which to release its heat. However,
the time a given amount of water
stays inside the fan-coil is irrelevant
in a system with continuous flow.
As previously discussed, the faster
Figure 4-19
equal air inlet temperature and air flow rates
a fluid moves along a surface, the
higher the rate of convective heat
transfer. Higher rates of convective
heat transfer between the water
and inside surface of the tubes
EQUAL RATES OF HEAT TRANSFER
<<180 ºF
1 row coil
180 ºF
2 row coil
<180 ºF
3 row coil
4 row coil
<<<180 ºF
equal discharge air temperature
decreasing entering water temperature
making up the coil result in higher
heat outputs.
Another way of justifying this principle
is to consider the average water
temperature in the coil at different
flow rates. As the flow rate through
the coil heat exchanger increases,
the temperature drop from inlet to
outlet decreases. This implies that
the average water temperature
within the coil increases, and so
does its heat output. Those who
claim that a smaller temperature
difference between the ingoing
and outgoing fluid implies less heat
is being released are overlooking
the fact that the rate of heat
transfer depends on both the
temperature difference and flow
rate. This relationship is described
by Formula 4-8 discussed earlier
in this section. This non-linear
relationship between flow rate and
heat output also applies to other
heat emitters such as radiant panel
circuits, panel radiators and finnedtube
baseboard. It can be seen
whenever heat output ratings are
plotted against flow rate through
the heat emitter. Figure 4-18 shows
this relationship for a small fan-coil
operated at a constant inlet water
temperature, constant air inlet
temperature and constant air-side
flow rate.
The rate of heat output increases
very quickly at low flow rates, but
less quickly as flow through the coil
increases. The fan-coil represented
in Figure 4-18 released about 50
percent of its maximum heat output
capacity at about 10 percent of its
maximum flow rate. The strong
curvature of this relationship implies
that attempting to control the heat
output of a fan-coil or air handler
by adjusting flow rate can be tricky.
Very small valve adjustments can
create large changes in heat output
at low flow rates. However, the
same amount of valve adjustment
will create almost no change in heat
49
output at higher flow rates. This is
also true for other types of hydronic
heat emitters.
Special types of balancing valves
known as “equal percentage” valves
have been developed specifically to
compensate for this characteristic.
An equal percentage valve allows
very small increases in flow rate
as the valve begins to open from
a closed position. The farther the
valve opens. the faster the rate at
which flow-through increases. The
combination of the heat output
versus flow rate characteristic of the
coil heat exchanger, and the flow
versus stem position characteristic of
an equal percentage valve, attempts
to make the change in heat output
approximately proportional to the
percentage of stem travel in the valve.
More information on equal
percentage valves can be
found in idronics 8 (Hydronic
Balancing).
The nonlinear relationship also
implies that attempting to boost heat
output from a fan-coil or air handler,
and other hydronic heat emitters,
by operating them at unusually high
flow rates will yield very minor gains.
The argument against doing this is
further supported by large increases
in pumping power with increasing
flow rate.
Principle #3: Increasing the air
flow rate across the coil heat
exchanger in a fan-coil or air
handler will marginally increase
the heat output of the fan-coil.
This principle is also based on
forced-convection heat transfer
between the exterior coil surfaces
and the air stream. Faster-moving
air reduces boundary layer thickness
and “scrubs” heat off the coil surface
at a higher rate. As with water flow
rates, the gain in heat output is very
minor above the nominal air flow rate
the unit is designed for.
Principle #4: Fan-coils with large
coil surfaces and/or multiple tube
passes through the fins can yield
a given rate of heat output while
operating at lower entering water
temperatures.
Conductive heat transfer from
the tubing carrying the fluid to
the aluminum fins of the coil heat
exchanger increases when more tubes
are used to make the coil. Convective
heat transfer is proportional to the
contact area between the surface and
the fluid. Both of these effects allow
larger liquid-to-air heat exchangers to
transfer a given rate of heat at lower
average water temperatures. This
principle is illustrated in Figure 4-19.
This principle is very important
for hydronic systems that operate
with low temperature heat sources
such as heat pumps. In these
applications, the designer may need
to use fan-coils with larger coils and/
or more tube passes through the
fins of the coil to drive heat from the
coil at the required rate. On larger air
handlers, manufacturers may offer
higher-performance coils with up to
eight tube rows.
Care must be taken that comfort is
not compromised when operating
fan-coils or air handlers at reduced
water temperatures. It is critically
important to introduce the air stream
from such units into the space so
that it mixes with room air before
passing by occupants. Failure to do
so will usually lead to complaints of
“cool air” blowing from the fan-coils
or registers, even though the room
is being maintained at the desired
temperature.
50
5. INSTALLATION DETAILS FOR HEAT EXCHANGERS
This Section discusses specific
details that allow heat exchangers to
perform as expected in several types
of applications.
HEAT EXCHANGER FOULING
To maintain peak performance, it’s
important to minimize the potential
for fouling films to form on any heat
transfer surfaces. Fouling can be
the result of unintentional chemical
interactions within the system. It can
also occur due to dirt or other debris
unintentionally present in the system.
CHEMICAL FOULING
Fouling films can form on heat
exchanger surfaces due to fluid
chemistry issues. For example, the
solubility of minerals such as calcium
and magnesium in water decreases as
water temperature increases. If water
containing these minerals passing
along the internal surfaces of a heat
exchanger becomes hot enough to
reach the maximum solubility limit
of the dissolved minerals, they will
precipitate out of solution and form
scale on those surfaces. The higher
the surface temperatures, the greater
the potential for scale formation.
Figure 5-1 shows an example of the
heat exchanger from a modulating/
condensing boiler that has been
severely scaled from minerals in the
system water.
The best approach for reducing
scaling of heat exchanger surfaces
is to demineralize the system water.
Demineralization is done by passing
water through a column containing
thousands of small porous polymer
beads that are chemically formulated
to capture positive and negative
ions in the system water. These
undesirable ions include positively
charged cations of calcium (Ca ++ ),
magnesia (Mg ++ ) and sodium (Na + ).
They also include negatively charged
anions of chlorine (Cl - ) and carbonate
(CO - - ). When the resin beads capture
these cations and anions, they release
hydrogen cations (H+) and hydroxide
anions (OH-). The released ions
instantly bond to form pure H 2 O.
Water in hydronic systems should
be deionized to a condition where its
total dissolved solids (TDS) content is
between 10 and 30 parts per million
(PPM). The small amount of remaining
ions gives the water sufficient electrical
conductivity to allow proper operation
of low-water cut-off devices. This level
of TDS also helps stabilize the water to
prevent it from scavenging ions from
metal surfaces.
idronics 18 provides more details
on demineralizing water in
hydronic systems
Figure 5-1
mineral scaling on hottest surfaces
scaling "flakes" accumulation
51
Figure 5-2
disassembled, cleaned, reassembled and put back in
service. In large systems, this could take several days and
will prevent heating or cooling operation in any portion of
the system involving the heat exchanger during that time.
Severely fouled brazed plate heat exchangers, or sealed
shell & coil heat exchangers, may need to be replaced.
Again, a costly and time-consuming process.
The best way to avoid these situations is to minimize the
possibility of debris entering heat exchangers. This can
be done by installing high-performance separators that
capture dirt and magnetic particles upstream of the heat
exchanger.
DEBRIS FOULING
Most hydronic systems are assembled from components
that have been stored, transported and handled multiple
times between manufacturing and installation. During these
times, debris such as road dust, insects, pollens, drywall
dust, solder balls, metal chips and sawdust can enter
these components. These materials can accumulate on
the internal surfaces of heat exchangers and significantly
reduce heat transfer rates. These accumulations can also
increase the pressure drop through components, resulting
in lower flow rates, which further reduce heat transfer.
Figure 5-2 shows a heavily fouled plate from a plate &
frame heat exchanger.
Although some of the heat exchangers discussed earlier in
this issue can be disassembled for cleaning, that process
is difficult, time consuming and expensive. The heat
exchanger must be isolated, disconnected from piping,
Some designers, by default, specify wye strainers as a
means of capturing dirt particles in the system. While this
approach is partially effective, it does have limitations. One
is that the pressure drop through a wye strainer increases
significantly as debris collects on the internal strainer
basket, as shown in Figure 5-3. The increased pressure
drop decreases flow in the circuit, which negatively impacts
the heat transfer performance of any heat exchanger.
Another limitation of wye strainers is that they must be
isolated and disassembled for cleaning. This requires
temporary shutdown of the circuit and some minor
drainage of system fluid.
Finally, wye strainers do not capture magnetic particles
such as magnetite that form when dissolved oxygen in
the water fluid reacts with steel or cast iron components in
the system. Magnetite will be attracted to magnetic fields
generated by circulator motors, especially the strong fields
generated by permanent magnets in wet rotor circulators
with electronically commutated motors.
Figure 5-3
∆P across strainer is
monitored to determine
when cleaning is necessary
basket strainer
ball valves to isolate
strainer during cleaning
pressure drop (psi)
6
5
4
3
2
1
0
0 2 4 6 8 10 12 14 16
flow rate (gpm)
1" Y-strainer
(70% plugged)
1" Y-strainer
(clean basket)
52
Figure 5-4a
Figure 5-5
isolation /
purge valve
plate
heat
exchanger
isolation flange
w/ purge valve
DirtMag
separator
DirtMag
separator
Figure 5-4b
isolation flange
w/ purge valve
isolation /
purge valve
Figure 5-6
isolation /
purge valve
closed
DirtMag
separator
closed
DirtMag
separator
The modern alternatives to wye
strainers are low-velocity zone
separators equipped with removable
magnets. Figure 5-4 shows examples
of Caleffi DIRTMAG separators
for both small and larger system
applications.
isolation flange
w/ purge valve
bucket with
cleaning solution
double
ended
hoses
Low-velocity zone dirt separators
do not allow the dirt they capture to
accumulate in the flow path, and thus
create far less pressure drop than wye
strainers. They can also be flushed
without shutting down the circuit.
submersible pump
53
Figure 5-7a
Figure 5-7b
DirtMag
separator
isolation /
purge valve
isolation flange
w/ purge valve
Some dirt separators are available
with strong permanent magnets
that can capture very fine iron oxide
particles formed from ferrous metals
in the system.
submersible pump
domestic water side
closed
closed
bucket with
cleaning solution
These DIRTMAG separators should
be installed upstream of heat
exchangers, as shown in Figure 5-5.
The component arrangement
shown in Figure 5-5 has the DirtMag
separators upstream of both the
heat exchanger and the circulators
supplying it. A combination isolation
flange/purge valve is shown on
the outlet of each circulator. These
flanges allow the circulator to be
isolated and removed if necessary.
In combination with isolation/purging
valves on the outlet ports of the heat
exchanger, they also allow each side
of the heat exchanger to be isolated
and flushed with a cleaning solution
if necessary. Figure 5-6 shows how
the heat exchanger could be flushed
by circulating a cleaning fluid using a
small submersible pump.
Figure 5-7 shows a slight variation on
valving, which can be used when one
side of the heat exchanger operates
with domestic water.
A combination isolation/purging valve,
rated for use with potable water, is
installed at each port on the domestic
water side of the heat exchanger. The
inline balls of these valves are closed
to isolate the heat exchanger from
the other domestic water piping. The
cleaning solution is circulated through
the domestic water side of the heat
exchanger to dissolve and remove
mineral scale. The two hoses can also
be reversed to change the direction
of the cleaning solution through the
exchanger. After the cleaning solution
has been sufficiently circulated and
drained, the heat exchanger should
be flushed with domestic water to
remove any residual cleaning solution.
HEAT EXCHANGERS SUPPLYING
ANTIFREEZE-PROTECTED CIRCUITS
When a heat exchanger is installed
as the “heat source” for a portion
of the overall system, that portion is
fully isolated from the remainder of
the system. As such, it must include
components used in all closed-loop
hydronic systems. These include
a pressure-relief valve, expansion
tank, pressure gauge, provisions
for filling and purging, and one or
more separators to remove air, dirt
and magnetic particles. Optional
components include a flow meter and
thermometers or temperature sensor
wells to monitor the inlet and outlet
temperatures of the heat exchanger.
In some applications, specifically
district heating systems, heat meters
are also installed on the closed circuit
served through the heat exchanger.
Figure 5-8 shows a configuration of
these components for a situation in
which the heat exchanger supplies
a garage floor heating system that
operates with an antifreeze solution.
The heat source is assumed to be a
convention boiler.
Whenever a conventional boiler is
supplying a low temperature heat
emitter system, that boiler should
54
Figure 5-8
temperature sensor
setpoint
controller
DiscalDirtMag
separator
SEP4
injection
circulator
PRV
slab on grade
floor heating
in garage
injection
mixing
controller
expansion
tank
antifreeze in this portion of system
conventional boiler
Figure 5-9
temperature sensor
setpoint
controller
thermostatic
mixing valve
Discal
ThermoProtec
valve
PRV
small
manifold
station
DirtMag
expansion
tank
antifreeze in this portion of system
conventional boiler
55
Figure 5-10
isolation flange
w/ purge valve
Discal
heat source
is heat pump
electric boiler
or mod/con
boiler
DiscalDirtMag
separator
PRV
setpoint
controller
DirtMag
expansion
tank
antifreeze in this portion of system
be protected against sustained flue gas condensation.
Consider that a garage floor slab that has not operated
for several days during the winter could be very cold —
even below 32ºF — when this portion of the system starts
operating. The very cold antifreeze solution returning from
the floor circuits poses two “threats” to the system. One is
that it could freeze the water on the primary side of the heat
exchanger. Given the rapid response of a brazed plate heat
exchanger, and its low water content, ice crystal can form
quickly — in some cases, in less than one minute. Any delay
in getting heat to the primary side of the heat exchanger
increases the possibility of freezing. A failure of the circulator
responsible for circulating heated water through the primary
side of the heat exchanger could also lead to freezing.
One way to protect against this situation is to install
a temperature setpoint controller that monitors the
temperature of water leaving the primary side of the heat
exchanger. When this temperature rises to some setpoint
deemed sufficient to verify that heat is being transferred to
the heat exchanger, this controller turns on the circulator
for the antifreeze side of the heat exchanger.
The other “threat” posed by a very cold slab is sustained
flue gas condensation in a conventional boiler. Keep in
mind that the rate of heat transfer to a very cold slab could
be significantly higher than under normal operation. Heat
exchangers will “transmit” this high rate of heat flow back
to the heat source.
One approach to protecting the boiler is to use an
injection mixing controller to regulate the temperature of
the antifreeze solution supplied to the floor circuits. This
controller also monitors the inlet temperature to the boiler
and reduces the flow of hot water to the primary side of the
heat exchanger, when necessary, to prevent sustained flue
gas condensation. A motorized 3-way mixing valve and
suitable controller is another option that can provide both
supply temperature control and boiler protection.
Another approach to this scenario uses two thermostatic
mixing valves, as shown in Figure 5-9.
The 3-way mixing valve on the right side of the heat
exchanger regulates the supply temperature to the floor
circuits. It is important to select this mixing valve with a Cv
rating close to the total flow rate required by the manifold
station. This will limit the pressure drop across the mixing
valve to approximately 1 psi. A Caleffi ThermoProtec valve
on the left side of the heat exchanger limits hot water
flow through the primary side of the heat exchanger, as
necessary, to keep the boiler inlet temperature high enough
to prevent sustained flue gas condensation.
When the heat source is a heat pump, electric boiler or mod/
con boiler, it is not necessary to using mixing devices in the
system. Figure 5-10 shows a representative configuration.
The heat exchanger should be selected so that it can
supply the design heating load to the floor circuits while
operating at an approach temperature drop of no more
than 5ºF. This allows the heat source to operate just a
few degrees F above the supply temperature of the floor
circuits. Low operating temperature improves the efficiency
of mod/con boilers and heat pumps.
56
6. HEAT EXCHANGER APPLICATIONS
This section describes several common applications, as
well as some “novel” applications for heat exchangers in
hydronic heating and cooling systems. They include:
Figure 6-1
• Snow & ice melting
• Pool heating
• Domestic water heating
• Heat input and extraction from non-pressurized
thermal storage tanks
• District heating
• Heat interface units (HIUs)
• Lake water cooling
Conceptual piping diagrams are presented for these
applications. Designers need to determine the suitability
of the concepts, as well as the required specifications of
the heat exchangers, piping, circulators, etc., for specific
applications. References to other issues of idronics are
given to provide further design assistance.
Courtesy of Gary Todd & REHAU
SNOW & ICE MELTING (SIM) SYSTEMS
The use of hydronics technology for snow and ice melting
dates back to the mid-1900s, when copper tubing was
embedded in exterior pavements. Today, PEX, PERT and
PEX-AL-PEX tubing provides a reliable way to distribute
heat from a boiler to exterior pavements such as sidewalks,
steps, driveways, emergency room arrival areas and even
complete parking lots.
SIM systems are categorized based on how they are
expected to perform relative to snow fall rates and
subsequent melting times.
Class 1 SIM systems are commonly used for residential
walkway and driveway areas. They deliver heat to
pavement at rates between 80 and 125 Btu/hr/ft 2 . These
rates are not necessarily high enough to melt snow as it
falls, especially during strong storms. Several hours may be
needed to completely melt the snow, especially when the
system starts from a cold condition.
Class 2 SIM systems are generally sufficient for pavements
at retail and commercial sites. They deliver heat at rates
between 125 and 250 Btu/hr/ft 2 . This class of SIM can
usually melt snow as it falls — once the pavement is up to
normal operating temperature.
Class 3 SIM systems are for the most demanding
applications where pavements must remain free of snow
and ice at all times. Examples include steeply sloped
exterior ramps at parking garages and hospital emergency
room entrances. Typical heat delivery rates range from 250
to 450 Btu/hr/ft 2 .
Courtesy of REHAU
57
Figure 6-2
boiler staging /
modulation controller
to / from
other loads
PRV
SEP4
hydraulic
separator
3-way
mixing valve
aquastat
or setpoint
controller
dual port
filling &
purge valve
Snow & ice melting
subsystem
antifreeze in this portion of system
ASSE 1017anti-scald
tempering valve
indirect water heater
(prioritized load)
Some SIM systems are single purpose. They are
designed solely for snow and ice melting. These systems
are usually supplied by one or more boilers. The entire
system operates with an antifreeze solution, and no
heat exchangers are needed between the boiler(s) and
distribution system.
In other systems, snow and ice melting is one of several
loads served by a common heat source. The other loads
are typically space heating and domestic water heating.
In these systems, a heat exchanger is often used to
separate the SIM subsystem from the remainder of the
system. This allows the majority of the system to operate
with water rather than antifreeze. Figure 6-2 shows a
representative configuration.
Figure 6-3 shows portions of snowmelting systems
supplied through plate & frame heat exchangers.
58
Figure 6-3a Figure 6-3b
Courtesy of West Coast Hydronics & Plumbing, Sacramento CA
POOL HEATING
Installations in which a boiler or hydronic heat pump supplies
space heating in winter and year-round domestic hot water
often have plenty of available heating capacity between the
end of one heating season and the beginning of the next. If
the property has a pool, that available heating capacity can
be used to maintain it at comfortable conditions. Figure 6-4
shows how a heat exchanger specifically designed for pool
heating can be installed.
This system is very similar to the one in Figure 6-2. The SIM
subsystem from Figure 6-2 has been replaced by a pool
heating subsystem.
Most heat exchangers intended for pool heating use a shell
& tube design, which is better suited to higher flow rates
needed for pool filter systems. Water from the heat source
passes through the shell of the heat exchanger, which can
be made of steel, cast iron or cast brass, and is part of
the closed-loop system. Pool water passes through the
tube bundle, which is typically made of titanium or titanium
alloyed with stainless steel to resist the corrosive effects of
highly chlorinated pool water.
The subsystem shown in Figure 6-4 uses the pool filter
pump to move pool water through the tube bundle in
the heat exchanger. Bypass piping is shown along with a
manually operated bypass valve. The shell side of the heat
exchanger is also piped to CPVC unions and ball valves.
This detailing allows the heat exchanger to be removed
from service if necessary, while still maintaining the ability
to circulate pool water through the filter.
Swimming pool water temperature is typically maintained
between 75 and 90ºF. This relatively low temperature allows
heat sources such as mod/con boilers and hydronic heat
pumps to operate at high efficiencies — assuming that the
heat exchanger is sized for a “tight” approach temperature
difference of perhaps 5ºF.
If one or more conventional boilers are used as the
heat source, it is important to protect the boiler(s) from
sustained flue gas condensation. This can be done using
Caleffi ThermoProtec valves on the boiler(s), as shown
in Figure 6-5.
Swimming pools have very high thermal mass. Without proper
detailing, this mass can “dominate” the water temperature in
the system, possibly preventing other subsystems that may
be operating at the same time from meeting their respective
loads. This condition must be avoided.
One approach is to use controls that make pool heating
a lower priority load relative to domestic water heating.
Thus, if the domestic water-heating load is active, flow to
the primary side of the pool heat exchanger is temporarily
suspended. This is essential if the boilers are programmed
59
Figure 6-4
boiler staging /
modulation controller
to / from
other loads
CPVC ball
valves
& unions
SEP4
hydraulic
separator
pool heat
exchanger
bypass valve
(normally
closed)
filter
pool
pool pump
ASSE 1017anti-scald
tempering valve
indirect water heater
(prioritized load)
60
Figure 6-5
to/from
space heating
load
boiler staging
controller
temperature sensors
variable-speed
circulator controlled by
temperature leaving
hydraulic separator
ThermoProtec
valve
ThermoProtec
valve
SEP4
hydraulic
separator
pool heat
exchanger
bypass
valve
(normally
closed)
filter
to/from
indirect water
heater
pool
pump
pool
to operate at higher water temperatures during the
domestic water-heating mode.
Another method is to use a “temperature-controlled”
variable-speed circulator to supply water to the primary
side of the pool heat exchanger. The speed of this circulator
is controlled based on the water temperature leaving the
hydraulic separator. If that temperature is below some
predetermined value — deemed necessary to allow proper
operation of other loads — the speed of the circulator
decreases. This reduces the rate of heat transfer to the
primary side of the pool heat exchanger. If the temperature
is above this setpoint, the circulator operates at full speed.
This detail is also shown in Figure 6-5.
DOMESTIC WATER HEATING
The most common method of heating domestic water in
hydronic systems that also supply space heating is by using
an indirect water heater. Figures 6-2 and 6-4 show how
such a water heater would be installed as a subsystem.
Indirect water heaters use internal coil heat exchangers.
The rate at which they can transfer heat to the domestic
water is typically limited by natural convection on the
outer surface of the coil heat exchanger. In applications
requiring very high rates of heat transfer, it may not be
possible to find an indirect water heat with sufficient heat
transfer capacity. In these situations, the use of a properly
sized brazed plate heat exchanger can meet the heat
transfer requirements. Figure 6-6 shows a possible piping
configuration for this approach.
Assume that the water in the upper portion of the tank needs
to be maintained at 140ºF. The stainless steel brazed plate
heat exchanger would be sized to transfer the full output
of the multiple boiler system with the water temperature
supplied from the boilers at 145ºF. Assuming a nominal
20ºF temperature drop through the primary side of the heat
exchanger, the 5ºF approach temperature difference keeps
the boilers shown in Figure 6-5 operating at inlet water
temperatures that produce some flue gas condensation,
and thus an incremental increase in efficiency.
61
Figure 6-6
boiler staging /
modulation controller
DHW
DHW riser
DHW riser
DHW riser
LegioMix
station
ThermoSetter
balancing
valves
recirculation
circulator
to/from
other
load
plate
heat
exchanger
SEP4
hydraulic
separator
check
valve
isolation / flush valves
(@ both potable water ports
of heat exchanger)
check
valve
stainless
steel
circulator
DHW storage tank
Combination isolation/flushing valves are installed on
each domestic water port of the heat exchanger. These
valves allow that side of the heat exchanger to be
periodically isolated and flushed with a cleaning solution
to remove scaling.
A stainless steel circulator is used between the heat
exchanger and storage tank. The circulator supplying
boiler water to the heat exchanger can be cast iron, since
it is part of the closed-loop system. Designers must ensure
that these circulators, as well as the associated piping, are
sized for flow rates that can transfer the full output of the
boiler system to the domestic water tank when necessary.
For example, if the maximum continuous domestic water
demand was 15 gpm, with a temperature rise from 50 to
140ºF, the required heat transfer rate is:
Assuming a 20ºF temperature change across the primary
side of the heat exchanger (e.g., 145ºF entering down to
125ºF leaving), the flow rate into the primary side must be
67.5 gpm. This will require a 2.5” copper tube to maintain
a reasonable flow velocity.
A commercially available sizing software suggests that
the minimum heat exchanger for this heat transfer
requirement is a 10” x 20” x 50 plate brazed plate
unit. At this flow rate, the pressure drop through that
heat exchanger is about 1.7 psi ( about 4 ft. of head).
Assuming that the other components and piping
between the hydraulic separator and heat exchanger
add 5 feet of head, the selected circulator needs to
provide a minimum of 67.5 gpm at 9 feet of head. This
will likely require a light commercial class circulator, not
a small zone circulator. If the wire-to-water efficiency of
that circulator is estimated at 30%, the estimated input
power to the circulator would be about 380 watts.
Unfortunately, there have been systems where an
undersized circulator, along with undersized piping, have
created flow “bottlenecks” in this approach to highcapacity
domestic water heating. Flow bottlenecks cause
62
Courtesy of Diversified Heat Transfer, Inc.
Figure 6-7
heat transfer bottlenecks. The latter
can severely limit the heat transfer
rate between the boilers and
domestic hot water. Under these
conditions, the boilers will quickly
reach their high limit settings and
short cycle. The boiler plant may
have sufficient heating capacity
for this high demand DHW load,
and the heat exchanger may have
sufficient heat transfer capacity, but
without adequate flow, the overall
process can be severely limited.
Pre-engineered and pre-assembled
products that use external brazed
plate heat exchangers in combination
with domestic hot water storage
tanks are available for high-volume
commercial applications. Figure 6-7
shows one example.
ON-DEMAND DOMESTIC
WATER HEATING
Another approach to domestic water
heating uses a brazed plate heat
exchanger sourced from a buffer
tank, or a boiler system, to create
domestic hot water literally seconds
before it is required at a fixture.
The assembly shown in Figure 6-8
can be used in residential and light
commercial systems that have a
heated buffer tank. That tank may
be associated with renewable heat
sources, such as solar thermal
collectors, biomass boilers or heat
pumps. It could also be heated by a
fossil fuel boiler.
When the domestic hot water drawn
by the plumbing distribution system
Figure 6-8
brazed plate heat exchanger
DiscalDirtMag
separator
isolation / flush valves
(@ both potable water ports of
heat exchanger)
PRV ASSE 1017
mixing valve
to / from heat source
check
valve
medium or
higher
temperature
water
CW
DHW
flow
switch
relay
circulator w/ check
& purging flange
to / from other loads
buffer tank
check
valve
63
Figure 6-9a Figure 6-9b
Courtesy of Harwill
exceeds 0.7 gpm, a flow switch, rated
for use with domestic water, closes it
contacts. This energizes the coil of a
relay, which turns on a circulator to
draw hot water from the top of the
buffer tank through the primary side
of the heat exchanger. After passing
through the heat exchanger, this flow
returns to the lower portion of the
buffer tank. Cold domestic water
flows through the secondary side of
the heat exchanger in a counterflow
direction, absorbing heat from the
primary side.
Figure 6-9a shows an example of
a flow switch suitable for use with
domestic water. The switch closes
its contacts at flow rates of 0.7 gpm
or higher and opens its contacts at
flow rates of 0.4 gpm or less. This
switch is designed to thread into a
3/4” stainless steel tee, as shown in
Figure 6-9b.
The high surface area-to-volume ratio
of the brazed plate heat exchanger,
combined with a minimal amount
of piping connecting it to the buffer
tank, allows heated domestic water
to be available within 3 to 5 seconds
of the flow switch contact closure.
This is significantly faster than the
response time required by a gasfired
tankless water heater. This
approach also has the advantage of
not storing domestic hot water when
it is not operational, which reduces
the potential for Legionella growth.
The system shown in Figure 6-8
assumes that the water in the buffer
tank is maintained hot enough to fully
heat the domestic water on a single
pass. This may or may not be the
case, depending on the heat source,
the space heating emitters, the
required domestic hot water delivery
temperature and the sizing of the
heat exchanger. When the thermal
storage tank is relatively hot, and the
heat exchanger is sized for a closeapproach
temperature difference,
there is no need for a temperature
boost after the domestic water leaves
the heat exchanger. An ASSE 1017
point-of-distribution mixing valve
should always be installed to protect
the fixtures from hot water delivery
temperatures over 120ºF.
This application is a good example
of when a heat exchanger with a
high thermal length, as discussed
in Section 4, is appropriate. The
domestic water needs to undergo
a large temperature change, but
the LMTD between the two fluids
will be relatively small. This type of
application will favor a brazed plate
heat exchanger with relatively long
plate length.
This configuration includes two
isolation/flushing valves installed at
both domestic water ports of the heat
exchanger. These valves allow the
heat exchanger to be isolated from
the rest of the domestic water system
and flushed with a cleaning solution,
as described in Section 5.
Another detail is the use of two
spring-check valves adjacent to the
buffer tank. The check valve near
the upper left connection prevents
reverse thermosiphoning between
the buffer tank and heat source. The
valve in the lower right limits heat
migration into the return side of the
space-heating circuits at times when
space heating is not required.
The brazed plate heat exchanger and
its associated piping should be kept
as close to the buffer tank as possible.
64
Figure 6-10
isolation / flush valves
(@ both potable water ports of
heat exchanger)
to / from heat source
check
valve
low to
medium
temperature
water
brazed
plate
heat
exch.
flow
switch
relay
electric
tankless
water
heater
DHW
CW
ASSE 1017
mixing valve
PRV
to
space
heating
buffer tank
check
valve
from
space
heating
This minimizes the amount of system
water that needs to be displaced at
the start of a DHW draw and helps
keep the response time to only a few
seconds. The heat exchanger and
its associated piping should also be
insulated to reduce heat loss to the
mechanical room.
Figure 6-10 shows an extension of
this approach. An electric tankless
water heater has been added to boost
the final DHW delivery temperature
when necessary. This configuration
would be useful when the buffer
tank operates at temperatures below
the required delivery temperature
for DHW. In this case, the heat
exchanger can still provide significant
preheating, leveraging the low
starting temperature of cold domestic
water. Combination isolation/flushing
valves are also installed on the piping
connected to the tankless water
heater to allow isolation and cleaning.
It would also be possible to use a
tank-type water heater to provide the
final temperature boost.
Another variation on this approach to
domestic water heating is shown in
Figure 6-11.
This system uses a brazed plate
heat exchanger to transfer heat
from a buffer tank to a recirculation
loop serving several DHW loads.
The system shown in Figure 6-11 is
based on use of a renewable energy
heat source, such as a heat pump,
to maintain the buffer tank in a
low- to medium-temperature range
— perhaps 90 to 110ºF. At these
temperatures, the heat pump would
operate at high COPs. If a spaceheating
load is also supplied by the
system, as shown in Figure 6-11,
the water temperature must be
compatible with the heat emitters.
Several DHW load points are
connected to the recirculation loop.
Each is equipped with a point-of-use
tankless electric water heater, which
would provide the temperature
boost required by each DHW
load. Each tankless heater can be
individually operated, and each can
have a different delivery temperature
setpoint.
65
Figure 6-11
to / from
space heating
point-of-use
instantaneous
water heaters
kitchen
bathroom group #1
bathroom group #2
Anti-Legionella
thermal sterilization
tankless heater
∆P valve
set for
1 to 1.5 psi
brazed
plate heat
exchanger
T&PV
DHW
CW
ASSE 1070
valve
renewable
energy
heat source
low to medium
temperature water
high efficiency
circulator
(continuous
oepration)
check
valve
SS
recirculation
circulator
insulated recirculation loop
3-way
diverter
valve
temperature
sensor in well
isolation &
flushing valves
CW
temperature
sensor (T1)
buffer tank
If the water in the recirculation loop reaches a temperature
equal to or above the DHW delivery temperature of one or
more of the tankless heaters, the heating elements in those
heaters do not operate.
The tankless heater and 3-way diverter valve at the
far end of the recirculation loop is solely for thermal
disinfection. Once each day, the diverter valve is powered
on to direct flow through this tankless heater, which has
a setpoint of 140ºF. This raises the temperature of the
recirculation loop and holds it there for at least 30 minutes
to kill any Legionella bacteria in the recirculation loop. Hot
water drawn from this loop during the disinfection cycle
passes through the ASSE 1070 point-of-use anti-scald
mixing valves, which limit temperature to the fixtures to
120ºF. After the thermal disinfection cycle is complete,
the dedicated tankless heater is powered off, and the
diverter valve directs flow around it. The tankless heater
used for thermal disinfection only needs to be sized to
the heat loss of the recirculation system when operating
at peak disinfection temperature. This may only be 3 or 4
kilowatts, depending on the amount of piping used and
how it is insulated.
It is important to insulate all portions of the recirculation
system, including the heat exchanger, to limit heat loss into
the building.
The brazed plate heat exchanger would be sized for a tight
approach temperature difference. It should be capable of
transferring the full heat output rate of the renewable energy
heat source when the tank is at its minimum temperature.
The buffer tank is configured in a “2-pipe” configuration. The
piping supplying warm water for low-temperature spaceheating
tees into the tank headers just upstream of the
tank. A differential pressure valve is used in the heat source
circuit to provide minimum forward opening pressure of 1
to 1.5 psi. This prevents water returning from the spaceheating
distribution system from passing through the heat
source when it is not operating. It also replaces the need for
a check valve to prevent reverse thermosiphoning between
the buffer tank and heat source.
If the renewable energy heat source supplying the buffer
tank is turned on and off automatically, as would be the
case with a heat pump or pellet boiler, that on/off operation
must be based on the temperature of the water in the tank.
The concept of on-demand domestic water heating using an
external heat exchanger can also be scaled up to high-volume
commercial applications. Figure 6-12 shows one example
of a pre-engineered, pre-assembled system that uses hightemperature
boiler water in combination with a plate & frame
heat exchanger for high-volume domestic water heating.
66
Courtesy of Diversified Heat Transfer, Inc.
Figure 6-12
The product shown in Figure 6-12 includes a modulating
control valve to control boiler water flow through the
primary side of the heat exchanger, and thus regulate
domestic water delivery temperature. It also has controls,
a stainless steel recirculation circulator, strainers and
pressure-relief valves.
A final thought on the use of brazed plate or plate &
frame heat exchangers for domestic water heating. Some
practitioners argue against this practice. The common retort
is that hard water can scale such heat exchangers. While
this is a possibility, it’s also a possibility with internal coil
heat exchangers, especially when the latter are operated
at high internal temperatures to force a given rate of heat
exchange through a relatively small surface area.
Most indirect water heaters have permanently installed
internal coil heat exchangers. When the exterior surface
of that coil is scaled, there is no practical way to clean it.
This is not the case with a brazed plate heat exchanger
equipped with the combination isolation/flushing valves
shown in several figures involving domestic water within
heat exchangers. If the scaling is beyond the point where
chemical cleaning can correct it, the brazed plate heat
exchanger can be replaced. When the domestic water side
of the internal coil heat exchanger is heavily scaled, the
entire indirect tank needs to be replaced, which is typically
much more costly.
Hard domestic water will scale any heat exchanger operating
at temperatures high enough to cause dissolved minerals
to precipitate out of that water. The hardness needs to be
reduced through softening, or the scaling problems will
persist, regardless of the type of heat exchanger used.
HEAT EXCHANGERS FOR NON-PRESSURIZED
THERMAL STORAGE TANKS
Some renewable energy heating systems, such as those
supplied by pellet-fired boilers or cordwood gasification
boilers, require several hundred gallons of thermal storage.
In some systems, that storage is provided by pressurerated,
ASME-listed steel tanks. These tanks are expensive,
and due to their size, the logistics of moving them into
place within a mechanical room can be challenging, and
sometimes even impossible. One solution is the use of a
“panelized,” non-pressurized tank that can be moved into
the mechanical space in pieces and then assembled.
Non-pressurized thermal storage tanks are vented to the
atmosphere. As such, they are “open” to absorption of
oxygen from the atmosphere. Although it is possible to
design a hydronic system that can connect directly with a
non-pressurized tank, there are several limitations to this
approach that are not present in a “closed” hydronic system.
One issue is that any steel or cast iron components
connected to the water in the non-pressurized tank would
be subject to severe corrosion due to the higher dissolved
oxygen content of the water. There are chemicals called
oxygen scavengers that can partially compensate for this
corrosion, but the concentration of these chemicals must
be monitored and maintained, which unfortunately doesn’t
always happen. This can cause premature failure of
expensive hardware, including boilers, circulators, radiators
and other components made of carbon steel or cast iron.
A better solution is to separate the water in a nonpressurized
thermal storage tank from both the steel boiler
and any ferrous metal in the distribution system. This can
be done using internal coil heat exchangers mounted within
the storage tank, as discussed in Section 2. It can also be
done with external brazed plate heat exchangers. Of these
choices, the latter provides several advantages including:
• Forced convection on both sides of the heat exchanger
for much higher convection coefficients, and the ability to
achieve the necessary heat transfer rates using smaller
heat exchanger surface area.
• The ability to accurately estimate the performance of the
heat exchanger. It is difficult to find detailed performance
models for internal coil heat exchangers operating over a
wide range of conditions.
67
Figure 6-13
passive heat dump
normally open
zone valve
no vents or other
devices that could admit
air should be placed
above the water level
vent
PRV
zone supplies
PRV
(P1A)
(P1B) (HX1)
air space
bullhead
tees
(HX2)
(P2)
zone returns
cordwood gasification boiler
• The ability to service or replace the
heat exchanger if ever necessary
without opening the thermal storage
tank.
• The possibility of using a single heat
exchanger for both heat input and
heat extraction from the tank.
The system shown in Figure 6-13 uses
two brazed plate heat exchangers:
One between the cordwood
gasification boiler and the thermal
storage tank, and the other between
the tank and the distribution system.
The boiler-to-tank heat exchanger
(HX1) is sized to transfer the boiler’s
maximum rate of heat output to the
tank water with only 5ºF approach
temperature difference between the
water entering the heat exchanger
from the boiler, and the water leaving
it for the tank.
Circulator (P1A) runs at a constant
speed whenever the boiler is enabled
to operate. A variable-speed stainless
steel circulator (P1B) provides flow
between the heat exchanger and
tank. The speed of this circulator
varies in response to the boiler
inlet temperature. At boiler inlet
temperatures below 130ºF, circulator
(P1B) is off. Its speed ramps up as
boiler inlet temperature climbs above
130ºF, reaching full speed at an inlet
temperature of 140ºF. This control
action prevents the boiler from
operating with sustained flue gas
condensation.
The pipe carrying hot water from heat
exchanger (HX1) passes through the
tank wall above the water line and
terminates in a bullhead tee. This
releases water horizontally and helps
preserve beneficial temperature
stratification within the tank.
Circulator (P1B) is mounted low
relative to the water level inside the
tank. This increases the net positive
suction head available to the circulator
and helps prevent cavitation. This
is important because the water at
the top of the tank is at zero gauge
pressure and can be very hot. Both of
these conditions decrease the margin
against cavitation.
The piping that’s slightly above the
water level in the tank will be under
slight sub-atmospheric pressure.
There should not be any air vents,
valves or other devices in this piping
that could potentially allow air to
enter the piping.
Purging flanges are installed on both
ports of circulator (P1B). These allow
water to be forced into the piping
to displace air when the system is
commissioned.
68
Figure 6-14
modulating /
condensing
boiler
differential
temperature
controller
supply
temperature
sensors
vent
PRV
variable-speed
pressure-regulated
circulator
zone
supplies
temperature
sensor (T1)
(HX2)
(P2)
variable-speed
temperature -
controlled
circulator
temp.
sensor
(T2)
SEP4
hydraulic
separator
low
temperature
heat emitters
bullhead
tees
differential
temperature
controller
vent
PRV
check
valves
mixing
valve
variable-speed
pressure-regulated
circulator
zone
supplies
temperature
sensor (T1)
(HX2)
temp.
sensor
(T2)
SEP4
hydraulic
separator
low
temperature
heat emitters
bullhead
tees
(P2)
ThermoProtec
valve
conventional boiler
69
The details between the tank and load side heat exchanger
are similar to those on the heat input side. Circulator (P2) is
also a stainless steel circulator.
A variable-speed pressure-regulated circulator provides
flow to the zoned distribution system.
Because there are two completely isolated closed-loop
portions within the overall system, two pressure relief valves,
two expansion tanks and two make-up water assemblies
are required. The expansion tanks can be relatively small
because they only service their respective closed-loop subsystems.
The expansion and contraction of water within the
tank simply changes the water level slightly.
This system requires two heat exchangers between the
heat source and the heat emitters. There must be some
temperature drop across each of these heat exchangers
to create the necessary rates of heat transfer. These
temperature drops should be kept small (suggested values
of not more than 5ºF). The lower the temperature drop
across heat exchanger (HX2), the lower the temperature
of the thermal storage tank can get while still being able to
satisfy the heat transfer requirements of the heat emitters.
This can prolong the discharge cycle of the tank and
reduce the number of boiler firings required.
Figure 6-14 shows two possibilities for adding an auxiliary
boiler or re-pipe an existing boiler to the system shown in
Figure 6-13.
The distribution system shown in Figure 6-14 serves
low-temperature heat emitters. If a conventional boiler
is used, it should be protected against sustained flue
gas condensation by a Caleffi ThermoProtec valve. This
high flow-capacity mixing valve automatically blends hot
water leaving the boiler with cool water returning from the
distribution system so that the water entering the boiler is at
or above some minimum temperature. A typical minimum
inlet temperature for a gas-fired boiler is 130ºF.
The speed of circulator (P2) is regulated to maintain a
desired supply water temperature to the heat emitters.
This detail also limits the supply water temperature to the
heat emitters if and when the thermal storage tank is at a
relatively high temperature.
When the thermal storage tank cannot contribute heat
to the system, circulator (P2) would be off. The boiler
circulator would be turned on whenever the boiler is
operating. Check valves are used to prevent flow through
the heat exchanger when circulator (P2) is off, or through
the boiler when it is off.
Figure 6-14 also shows a differential temperature controller
in both systems. This controller is energized whenever one
or more zones are operating. It monitors the temperature
difference between the water near the top of the thermal
storage tank at sensor (T1), and the water returning from
the distribution system at sensor (T2). Circulator (P2) is
allowed to operate when one or more zones is operating,
and the temperature at sensor (T1) is 8ºF or more above the
temperature at sensor (T2). When this temperature difference
drops to 5ºF or less, circulator (P2) is turned off. This control
action prevents heat created by the auxiliary boiler from being
inadvertently transferred into the thermal storage tank. The
on and off settings for the differential temperature controller
can be adjusted based on the operating characteristics
of the distribution system and the approach temperature
difference across heat exchanger (HX2).
DISTRICT HEATING
There are situations where it’s desirable to heat several
buildings from a central heat plant, rather than use
individual heat systems within each building. Examples
include college campuses, multiple apartment buildings in
close proximity or groups of retail stores forming a shopping
center. This can be done using a district heating system.
Most district heating systems have a central boiler plant
where all heat is generated. That heat could come from
fossil fuel boilers, renewable heat sources such as biomass
boilers or a combination of these heat sources. The boiler
plant typically is configured as a multiple boiler system to
provide capacity control and provide redundancy if one
boiler is down for service.
Heat from the boiler plant is distributed to multiple “client”
buildings through insulated underground piping. The layout
of this distribution piping depends on the locations of the
client buildings. When the buildings are located along a street,
a 2-pipe direct return piping system is typically used and
installed parallel with the street. When the buildings form a
cluster around the boiler plant, a homerun piping distribution
system is appropriate. Figure 6-15 shows an example of the
latter, with one of three client buildings represented.
In this system, heat is generated by three modulating/
condensing boilers. Assuming that each boiler has a 10:1
turndown ratio, this provides a wide range of heat output
capability, specifically a 30:1 system turndown ratio. This
helps maintain good efficiency under partial load conditions.
The boilers supply hot water to the SEP4 hydraulic
separator, which provides air, dirt and magnetic particle
separation. It also isolates the pressure dynamics of the
boiler circulators from the variable-speed circulators that
provide flow to the client buildings.
70
Figure 6-15
multiple boiler
staging &
modultion
controller
CONTECA
heat meter
to other
buildings
SEP4
hydraulic
separator
insulated underground piping
(P1)
(S1)
BOILER PLANT
CLIENT BUILDING
make up water
variable-speed
pressure-regulated
circulator
CONTECA
heat meter
zone supplies
(P3)
(S2)
plate & frame
heat exchanger
PRV
zone returns
SEP4
hydraulic
separator
(P2)
make up water
isolation / flushing valves
Circulator (P1) is one of the distribution circulators. Its
speed is based on temperature feedback from sensor
(S2) on the supply side of the client building’s distribution
system. The controls for the system strive to maintain a
target temperature at sensor (S2). That temperature could
be a setpoint or based on outdoor reset control. If the
measured temperature at sensor (S2) is below the target
value, the speed of circulator (P1) is increased, and vice
versa. This control action reduces pumping energy to the
minimum needed for the current load in the client building.
71
Figure 6-16
to / from
other cooling
loads
variable-speed
pressure-regulated
circulator
zone supplies
(P3)
CLIENT BUILDING
plate & frame
heat exchanger
CONTECA
heat meter
to / from
district system
PRV
zone returns
(P2)
make up water
isolation / flushing valves
SEP4 hydraulic separator
ThermoProtec valve
existing boiler
A plate & frame heat exchanger
is used to isolate the boiler plant
and underground piping from the
distribution system in the client
building. This heat exchanger is fully
insulated to minimize heat loss. This
prevents a potential leak or servicing
requirement on the client side of the
heat exchanger from disrupting service
to the other client buildings. This heat
exchanger is provided with valves for
isolation and flushing if necessary.
The left side of the heat exchanger,
and all components connected to
that side, form a closed-loop system
that is completely separated from the
boiler plant and underground piping.
As such, this portion of the system
requires a pressure relief valve,
expansion tank, make-up water
assembly, circulator and provisions
for air, dirt and magnetic particle
separation. These functions are
provided by Caleffi SEP4 hydraulic
separators.
Two CONTECA heat-metering
systems are shown in this system.
The one in the boiler plant monitors
all heat leaving the plant to the client
building. The heat- metering system
in the client building monitors all heat
arriving at the building from the district
system. The difference in readings
between these two heat-metering
systems would give the thermal loss in
the underground piping. Some district
systems may only need one of these
heat-metering systems, depending
on how the thermal losses of the
underground piping will factor into the
energy charges for the client buildings.
For more information on heat
metering, see idronics 24.
Flow in the client building’s distribution
system is provided by a variable-speed
pressure-regulated circulator (P2). The
speed of this circulator automatically
adjusts as the zone valves in the client
building open and close.
There are many possible variations
on this system. For example, one or
more of the client buildings may have
72
Figure 6-17
an existing boiler that could provide heat to the building if
the district system was down. One piping design that could
accommodate this existing boiler is shown in Figure 6-16.
Zoning within the client building might use multiple
circulators rather than zone valves. Some district systems
can also be designed to provide chilled water cooling in
summer as well has heating in winter.
HEAT INTERFACE UNITS
Another application that relies on brazed plate heat
exchangers uses centralized heat production, or district
heating, to supply space heating and domestic hot
water for a multi-unit building containing apartments,
condominiums, office spaces or retail shops. Heat from a
central “plant” or district system is transported throughout
the building using a 2-pipe distribution system. A third pipe
carrying cold domestic water is routed along with these
two pipes. At each building unit, the three pipes connect to
a compact assembly called a heat interface unit (HIU). The
heat passing from the 2-pipe building distribution system
into each HUI is accurately metered. Each unit within the
building is then charged for the thermal energy they use.
The HIU also meters the total domestic water used.
Heat exchangers are a critical component within HIUs.
They allow the primary water supplied from the building’s
central heat plant to be separated from the water used for
space heating or domestic hot water within each unit. The
separation of heating plant water from domestic water is
an obvious requirement. The separation of heating plant
water from the water used for space heating within each
unit ensures that a potential leak in the space-heating
subsystem within that unit can be isolated and repaired
without disrupting service to other units served by the
central heating plant.
Figure 6-17 shows an example of a modern HIU that uses
two brazed plate heat exchangers — one for space heating
and the other for heating domestic water.
The HIU is mounted within a wall cavity in the building unit,
close to the distribution pipes from the central plant. It has
connections for the heating supply and return pipes, as
well as one connection for cold domestic water.
When space heating is required, a modulating valve within
the HIU regulates the flow of hot water from the central
plant through the primary side of one brazed plate heat
exchanger. This flow regulation is based on maintaining a
desired supply water temperature leaving the secondary
side of the heat exchanger and supplying the unit’s spaceheating
distribution system.
A flow switch detects when there is a demand for domestic
hot water. A second modulating valve then regulates flow
from the central plant through the primary side of a second
brazed plate heat exchanger. The flow regulation is based
73
Figure 6-18
To/from central
heating plant
Modulating valves
temp. sensors
Heating
interface
unit
To/from
space heating
in unit
Cold domestic
water main
on providing a desired domestic
hot water temperature leaving the
secondary side of the heat exchanger.
The total space heating and domestic
hot water energy used is based on
the flow rate and temperature drop
of the central plant water from where
it enters and leaves the HIU. The
controller within the HIU records the
total thermal energy used.
For more information on heat
metering, see idronics 24.
The HIU shown in Figure 6-17 also
includes a circulator and expansion
tank for the space-heating distribution
system within the building unit.
Exp.
tank
CW
DHW
Controller &
heat meter
Domestic
water
meter
DHW
flow
switch
Figure 6-18 shows a conceptual
piping arrangement for an HIU that
provides the functions just described.
It does not show exact hardware
devices or their placement.
HIUs allow individually metered
thermal energy for space heating
and domestic hot water without the
need for combustion heat sources
within each unit. This reduces cost
and increases safety by eliminating
the need for gas piping and venting
of combustion appliances dispersed
throughout a building.
LAKE WATER COOLING
At depths of approximately 40 feet or
more, the water temperature in lakes,
located in climates with cold winters,
remains relatively constant at about
39ºF. Water at this temperature is
ideal for chilled-water cooling, if it can
be accessed from shore.
The potential to use direct lake source
cooling for smaller buildings depends
on several factors. First, the building
must be located close to a lake, and
such that the intake pipe can be
located at a sufficient depth to extract
cool water. Ideally, the intake pipe
would be at least 30 feet beneath the
lake surface. However, lesser depths
may be possible depending on the
water temperatures required by the
cooling system.
Second, the project must conform to
any regulations regarding use of lake
water. Different requirements may
apply to “navigable waters” versus
lakes in which boating is not allowed.
Compliance checks should be done
at the onset of a feasibility study to
rule out possible restrictions that
would prevent further consideration
of this approach.
If regulations allow further pursuit
of the project, water temperature
measurements should be taken
at several depths below the lake
surface during late summer to assess
available water temperatures during
months when cooling is needed. This
will help in determining the required
depth of the water intake pipe.
Assuming the project reaches the
design phase, Figure 6-19 shows
one possible method for “harvesting”
cool lake water for building cooling.
This system uses readily available
components.
A shallow well pump provides flow
for the lake piping. This pump
might be within the building being
cooled or a separate building closer
to the lake. If water remains in the
system during winter, the pump and
piping must be protected against
freezing. Shallow well pumps are
limited in their ability to lift water
above the surface of the lake. If the
lift requirement is not more than
approximately 20 feet above the
lake surface, and the piping to and
74
Figure 6-19
piping
below
frost line
return pipe
lake
chilled water
air handlers
concrete pedestal
foot valve
variable-speed
∆P circulator
isolation & flush
valves
expansion
compensators
to / from
other cooling
loads
Plate & frame
heat exchanger
automatic
filter
SS pump
flush water
(30 gpm @ 30 psi min.)
check valves
drainage
from the lake is sized for low head
loss, a shallow well pump should
suffice. The objective is to keep the
pump as low as possible, relative to
the lake surface.
Lake water is drawn into a foot
valve at the end of a high-density
polyethylene (HDPE) pipe. The
end of this pipe and foot valve are
secured to a concrete ballast block
that prevents the pipe from floating.
This ballast block also suspends the
foot valve above silt on the lakebed.
Depending on the length of pipe
required, it may be necessary to use
additional concrete ballast blocks to
ensure the piping in the lake stays on,
or close to, the lakebed. The piping
must make the transition to shore
at depths that prevent the water it
contains from freezing in winter.
The incoming lake water passes
into an automatic filter. This device
monitors the differential pressure
across its inlet and outlet ports.
When the pressure differential
reaches a set value, the filter initiates
an automatic cleaning cycle. In this
mode, a typical automatic filter needs
a minimum flow of 30 gpm at 30 psi
pressure to provide proper cleaning.
This flow is provided from a separate
water source routed through a pilotoperated
solenoid valve that opens
simultaneously with the waste
valve on the filter. The accumulated
sediment is washed off the internal
stainless steel screen and blown
out through a waste pipe. A
filtering criterion of 100 microns
is suggested to minimize any
accumulation of sediment within
the heat exchanger. The frequency
of filter operation depends on the
characteristics of the lake. If the
intake pipe is located deep within a
lake that has minimal surface water
flow, the intake water should remain
relatively clean. However, if the
lake water experiences significant
weather-related disturbances, the
water may be turbid at times, and
thus will require more filtering.
Water passes from the automatic
filter to the pump. Ideally, this
pump should be constructed with
a stainless steel volute to provide
maximum corrosion resistance. Full
port ball valves should be installed
to isolate the pump if maintenance is
required. A priming line may also be
required to add water to the intake
pipe when the system is being put
into operation.
75
From the pump, lake water passes
into one side of a stainless steel plate
& frame heat exchanger. The heat
exchanger should be sized so that
the approach temperature difference
between the incoming lake water and
the chilled water leaving the other
side of the heat exchanger is not
more than 5ºF under design cooling
load conditions. Lower temperature
rises across the heat exchanger may
allow for smaller coils in air handlers,
lower flow rates and lower circulator
operating costs.
After passing through the heat
exchanger, the lake water returns to
the lake through another HDPE pipe.
Like the supply pipe, the return pipe
should be kept on or near the lakebed
using concrete ballast blocks. It
should discharge into the lake several
feet away from the foot valve, and
preferably in a direction that carries
the water away from the foot valve.
The load side of the heat exchanger
connects to a 2-pipe chilled-water
distribution system. The ECM-based
circulator operates on differential
pressure control and changes speed
in response to the opening and
closing of the zone valves controlling
flow to each air handler.
The electrical energy used by the
lake water pump could be further
reduced by using a variable-speed
circulator that responds to either
the temperature rise across the lake
side of the heat exchanger or to the
temperature of the lake water leaving
the heat exchanger. Either would
allow the flow rate of lake water to
be adjusted based on the current
cooling load on the other side of the
heat exchanger.
The lake water circuit could also be
used to supply low-temperature heat
to a water-source heat pump for
space heating or for domestic water
heating.
Figure 6-20
Figure 6-21
LAKE PLATE HEAT EXCHANGERS
Another method for harvesting the
cooling potential of lakes uses a
closed plate-type heat exchanger, an
example of which is shown in Figures
6-20 and 6-21.
The lake heat exchanger consists
of multiple stainless steel plate
assemblies. Each assembly has
two stainless steel plates that are
specially patterned to create flow
channels. These plates are welded
together along their perimeter. Each
Courtesy of AWEB Supply. Courtesy of AWEB Supply.
76
Figure 6-22
air handlers
balancing
zone valve
valves
variable-speed
∆P circulator
to / from
other cooling
loads
SEP4 Hydro
Separator
expansion
compensators
Lake plate heat exchanger
fill & purge
valves
silt
plate assembly is connected to a
supply and return header. The overall
assembly is welded to a stainless
steel base that supports the plates
several inches above the surface they
rest on. HDPE tubing is routed from
the headers to the shore.
This type of heat exchanger is
typically used with water source heat
pumps. However, when properly
sized and used in a lake where the
water temperatures at the lakebed
are stable and low, it could provide
direct heat exchange to a chilledwater
cooling system, as shown in
Figure 6-22.
This type of heat exchanger allows
for a completely closed hydronic
cooling system. This eliminates the
need for the automatic blowdown
filter, as well as the flat plate
heat exchanger. However, a dirtseparating
device designed for a
closed hydronic circuit should still be
installed to establish and maintain a
very low level of suspended solids
in the recirculating water. In Figure
6-22, a SEP4 hydraulic separator
provides air, dirt and magnetic
particle separation. It also allows for
different flow rates between the lake
heat exchanger and the chilledwater
distribution system.
Manufacturers of lake water
heat exchangers provide design
assistance software that can be used
to select a specific heat exchanger
based on total cooling load, lake
water temperature, flow rate through
the heat exchanger and temperature
change across the heat exchanger.
As with direct lake water cooling,
designers should verify if any
regulations prevent the placement
of a lake plate heat exchanger,
especially in navigable waters, as an
initial step in the feasibility of this type
of cooling system.
SUMMARY:
Heat exchangers add exceptional
versatility to modern hydronic
systems. They allow physical
separation of fluids, but passage of
heat, enabling applications such as
domestic water heating, snowmelting
and pool heating.
Section 4 gives the basis for
theoretical performance assessment
of heat exchangers. The associated
mathematics can be challenging.
Most practicing designers now use
software available from several heat
exchanger manufacturers to narrow
the design down to an appropriate
heat exchanger model.
To ensure reliable performance, heat
exchangers need to be maintained
with little if any fouling. The use of
demineralized water as well as highefficiency
dirt separators is important.
For more information on
demineralized water,
see idronics 18.
77
APPENDIX A: CALEFFI HYDRONIC COMPONENTS
CALEFFI HYDRONIC COMPONENTS
5517
Discal
rotating
collar
551
Discal
551
Discal
Compact
NA551 Discal
ASME / CRN
NA551 Discal
ASME / CRN
5020
Minical
5023
Valcal
501
Maxcal
553
AutoFill
5350
AutoFill
573
AutoFill
Combo
573 Dual Check
Backflow Preventer
574 RPZ
Backflow
Preventer
574
AutoFill
Combo
NA5453 DirtMag
rotating Collar 5463
DirtMag
NA5465 Dirtcal
ASME / CRN
NA5465M DirtMag
ASME / CRN
NA5465 Dirtcal
ASME / CRN
NA5465M DirtMag
ASME / CRN
Z2 Z-one
2-way
Z3 Z-one
3-way
6762 TwistTop
Zone Valve
574
AutoFill
Combo
6442 2-way
motorized
valve
6443 3-way
motorized
valve
280
ThermoProtec
521 MixCal
w/ Gauge
NA546
DiscalDirt
ASME / CRN
546 5461
DiscalDirt DiscalDirtMag
NA546M
DiscalDirtMag
ASME / CRN
5461
DiscalDirtMag
NA546 DiscalDirt
ASME / CRN
NA546M DiscalDirtMag
ASME / CRN
221
Radiator
Valve
3
220
Radiator
Valve
3
472 Thermostatic
Control Head
142
Flo-Set 130
Flo-Set
127 FlowCal 127 FlowCal+ 121 FlowCal
132 132
QuickSetter QuickSetter
519 Diff Pressure
Bypass Valve
NA51
Check Valve
120 Y-strainer
548
Hydro
Separator
5495
SEP4
NA548
Hydro Separator
ASME / CRN
NA549
SEP4
ASME / CRN
NA548 Hydro Separator
ASME / CRN
NA549 SEP4
ASME / CRN
663 Distribution Manifold
663 Distribution Manifold
Inverted
668S1 TwistFlow Manifold Assembly
172 Manifold Mixing Station
7504 Conteca Energy Meter
110 GeoCal w/ QuickSetter
7504 Conteca Energy Meter
sensor wells
7504 Conteca Energy Meter
flow meter
5599 HydroLink 5599 HydroLink
281 ThermoProtec
165 HydroMixer
166 HydroMixer
w/ Mix Valve
167 HydroMixer
w/ Elec Mix Valve
5599 HydroLink
5599 HydroLink
78
APPENDIX B: CALEFFI PLUMBING COMPONENTS
521 MixCal 520 AngleMix
w/ Gauge w/ Gauge
521 MixCal 520 AngleMix
w/ Gauge w/ Gauge
520 TankMixer
520 TankMixer
5213
521 MixCal 5212 Point-of-Use
5213
w/ Checks
521 MixCal SinkMixer Mix Valve
5212 Point-of-Use
w/ Checks SinkMixer Mix Valve
5231
MixCal
5231
MixCal
5231
MixCal+
5231
MixCal+
6000 LegioMix 6000 LegioMix
6000 LegioMix 6000 LegioMix
6000 LegioMix
Control
6000 LegioMix
Control
127 FlowCal 127 FlowCal+
127 FlowCal 127 FlowCal+
130
Flo-Set
130
Flo-Set
142
Flo-Set
142
Flo-Set
132 QuickSetter+
132 QuickSetter+
NA5026
PlumbVent
NA5026
PlumbVent
574 RPZ
Backflow
574 RPZ
Preventer
Backflow
Preventer
533H
PresCal
533H
PresCal
535H
PresCal
535H
PresCal
Z2 Z-one
536A 2-way
Z2 Z-one
PresCal HP Low Lead
536A 2-way
PresCal HP Low Lead
Z3 Z-one
3-way
Z3 Z-one
Low Lead
3-way
Low Lead
116 ThermoSetter
w/ Check
116 ThermoSetter 116 ThermoSetter
& Disinfection
w/ Check W/ Check
116 ThermoSetter
& Disinfection
NA51
W/ Check
Check Valve
NA51
Check Valve
6000
LegioMix
6000
Station
LegioMix
Station
APPENDIX C: GENERIC COMPONENTS
circulator w/
isolation flanges
circulator w/
isolation flanges
circulator w/
internal check
circulator w/
valve
internal check
& isolation
valve
flanges
& isolation
flanges
variable-speed
pressure-regulated
variable-speed
circulator
pressure-regulated
circulator
ball valve
ball valve
variable-speed
pressure-regulated
variable-speed
circulator w/ check
pressure-regulated
circulator w/ check
swing check valve
swing check valve
spring-loaded
check valve
spring-loaded
check valve
purging valve
purging valve
high capacity
circulator
high capacity
circulator
3-way motorized
mixing valve
3-way motorized
mixing valve
4-way motorized
mixing valve
4-way motorized
mixing valve
union
union
pressure
gauge
pressure
gauge
closely-spaced tees
closely-spaced tees
diverter tee
diverter tee
pressure
relief valve
pressure
relief valve
pressure &
temperature
pressure &
relief valve
temperature
relief valve
panel radiator
w/ isolation valve
panel radiator
& TRV
w/ isolation valve
& TRV
conventional boiler
conventional boiler
reversible water-to-water
heat pump
reversible water-to-water
heat pump
heating heating mode mode
reversing
valve
reversing
valve
evaporator evaporator
condenser condenser
TXV
TXV
modulating / condensing boiler
modulating / condensing boiler
firetube
mod/con
boiler
firetube
mod/con
boiler
monobloc air-to-water heat pump
monobloc air-to-water heat pump
solr thermal collector (array)
solr thermal collector (array)
split-system air-to-water heat pump
split-system air-to-water heat pump
(P4)
(P4)
hose bib
drain valve
hose bib
drain valve
diaphragm
expansion
diaphragm
tank
expansion
tank
horizontal
air handler
horizontal
air handler
brazed-plate
heat exchanger
brazed-plate
heat exchanger
plate & frame
heat exchanger
plate & frame
heat exchanger
indirect water heater
indirect water heater
wood-fired boiler
wood-fired boiler
pellet boiler
pellet boiler
79
www.caleffi.com
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